Method of and system for purifying exhaust gas for engines

Exhaust gas temperature is raised after a cold engine start so as to rapidly activate a catalyst in an exhaust line by retarding an ignition timing at which an air-fuel mixture is fired in the engine to a point after top-dead-center during a certain time interval and promoting combustion of the air-fuel mixture, thereby accelerating the catalyst to attain its effective activation temperature after the cold engine start as quickly as possibly.

BACKGROUND OF THE INVENTION 
1. Field of the Invention 
The present invention relates to a method of purifying the exhaust gasses 
for a reciprocating type of an internal combustion engine and an system 
for implementing the method. 
2. Description of Related Art 
Typically, catalytic converters, conventionally employed in exhaust systems 
for automobile engines, have the purpose of controlling exhaust gas 
emissions through the chemical conversion of hydrocarbons (HC), carbon 
monoxide (CO) and nitrogen oxides (NOx) into non-polluting components, 
such as water, carbon dioxide, and nitrogen and purifying the exhaust gas. 
Conventional three-way catalytic converters make use of precious metal 
catalysts, such as platinum, rhodium and palladium, incorporated into a 
silicon oxide carrier, and must operate above a specific activation 
temperature of, for example, 350.degree. C. in order to adequately convert 
exhaust gas emissions to the aforementioned non-polluting components. 
Because the temperature of exhaust gas, under normal engine operating 
conditions when the engine is in a warmed up condition, normally ranges 
from 700 to 900.degree. C. and the catalytic converter is kept at a 
temperature well above its activation temperature, the desired level of 
exhaust gas emission control can be adequately obtained. 
The catalytic converter, however, gradually increases in temperature with 
the exhaust gas when the engine operates before being sufficiently warmed 
up, the result being that exhaust gas emission control is ineffective for 
approximately the first 60 seconds after an engine start. As a result, 
hydrocarbons (HC), carbon monoxide (CO) and nitrogen oxides (NOx) 
contained in the exhaust gas are released into the atmosphere during the 
cold start until the catalytic converter reaches the effective activation 
temperature. This demonstrates the necessity of a means to quickly raise 
the temperature of catalyst to the effective activation temperature 
within, for example, approximately the first 20 seconds of engine 
operation after a cold start. 
Various means of dealing with this problem have been proposed. For example, 
Japanese Unexamined Patent Publication No. 4-66715 proposes a system or 
device of solving the problem through the installation of an electric 
catalyst heater in an exhaust duct upstream from the catalytic converter, 
or in the catalytic converter itself in close proximity to the inlet 
thereof, with the purpose of quickly raising the temperature of the 
exhaust gas so as to force the catalyst to reach the effective activation 
temperature while the engine is warming up. Otherwise, Japanese Unexamined 
Patent Publication No. 6-1672 proposes an afterburner device which is 
installed in an exhaust duct upstream from the catalytic converter, or in 
the catalytic converter itself in close proximity to the inlet thereof, 
with the purpose of quickly raising the temperature of the exhaust gas so 
as to force the catalyst to reach the effective activation temperature 
while the engine is warming up. 
These exhaust gas purifying systems equipped with the electric catalyst 
heater or the afterburner device are disadvantageous because these heater 
and afterburner demand construction of additional components which make 
the exhaust gas purifying system larger and more complex, and because they 
pose problems in regard to their integration into the overall design of 
the vehicle. Moreover, not only the electric heater demands a large amount 
of electrical power which necessitates the use of a higher capacity 
battery and alternator than would normally be required in the vehicle, but 
also a control box would be required for the electric heater, along with 
heavy gauge wiring, both of which would pose additional problems in 
integrating the design into the vehicle. 
SUMMARY OF THE INVENTION 
It is an object of the present invention to provide a method of and a 
system for exhaust gas purification which can quickly raise the 
temperature of catalyst to an effective activation temperature during 
engine warm-up after a cold engine start. 
It is another object of the present invention to provide an exhaust gas 
purifying system of simple construction which can be easily integrated 
into vehicle design. 
The aforesaid objects of the present invention are accomplished by 
providing a method of and a system for purifying exhaust gas from a 
reciprocating internal combustion engine of a vehicle which is equipped 
with an exhaust system having a catalytic converter. A rise in exhaust gas 
temperature to an effective activation temperature of the catalyst is 
controlled during a time interval when the engine starts in a cold 
condition, by means of exhaust gas temperature control in which engine 
operation is controlled so as to retard the ignition timing at which an 
air-fuel mixture is fired in the engine to a point after top-dead-center 
during the time interval for a necessary rise in exhaust gas temperature 
and to promote combustion of the air-fuel mixture, thereby accelerating 
the necessary rise in exhaust gas temperature so that the catalyst attains 
the effective activation temperature as soon as possible. The retardation 
of the ignition timing may be made according to a warmed condition 
detected at said engine start by said thermal condition monitoring means. 
According to an aspect of the present invention, the exhaust gas 
temperature control alters an in-cylinder air-fuel ratio at which the 
air-fuel mixture is introduced into the engine in a range from 13.5 to 
18.0 during the time interval for the rise in exhaust gas temperature. 
According to another aspect of the present invention, the exhaust gas 
temperature control alters an excessive air ratio of the air-fuel mixture 
greater than one (1) during the time interval for the rise in exhaust gas 
temperature. 
According to another aspect of the present invention, the exhaust gas 
temperature control alters the in-cylinder air-fuel ratio and supply part 
of intake air directly to the exhaust gas upstream from the catalytic 
converter so as to alter a virtual exhaust air-fuel ratio, which is 
defined as a ratio of a total amount of intake air including intake air 
introduced into the air-fuel mixture and the intake air supplied directly 
to the exhaust gas relative to an amount of fuel introduced into the 
air-fuel mixture, is at greater than 14.5. Regulating the amount of intake 
air independently from operation of an engine throttle valve according to 
a warmed condition detected at an engine start changes the in-cylinder 
air-fuel ratio and maintains a specified speed of the engine in the time 
interval for the necessary rise in exhaust gas temperature. 
According to another aspect of the present invention, altering the ignition 
timing to the point after top-dead-center is made after the engine has 
attained the specified speed of revolution subsequently to the cold start. 
The in-cylinder air-fuel ratio is feedback controlled based on a 
concentration of oxygen (O.sub.2) in the exhaust gas which may be 
monitored by a heater incorporated lambda oxygen (.lambda.O.sub.2) sensor 
so as to reach a stoichiometric air-fuel ratio as a target ratio during 
the time interval for the rise in exhaust gas temperature. 
According to another aspect of the present invention, the amount of intake 
air introduced into the engine independently from operation of the 
throttle valve is regulated at the beginning of the time interval for the 
rise in exhaust gas temperature so as to maintain the specified engine 
speed and thereafter the ignition timing is further retarded so as to 
decrease the engine speed below the specified engine speed during the time 
interval for the rise in exhaust gas temperature. 
According to another aspect of the present invention, the ignition timing 
is retarded to a point more later but before top-dead-center when the 
vehicle starts under a driving condition during the time interval for the 
rise in exhaust gas temperature than when driving under the same driving 
condition after the time interval for the rise in exhaust gas temperature. 
According to another aspect of the present invention, while the ignition 
timing is retarded to a point after top-dead-center during the time 
interval for a necessary rise in exhaust gas temperature to the effective 
activation temperature of the catalyst, enhanced turbulence, such as a 
swirl and/or a tumble, of the air-fuel mixture is generated in the 
cylinder so as to promote combustion of the air-fuel mixture. In order to 
generate enhanced turbulence of the air-fuel mixture acceleration of a 
flow of intake air entering the cylinder, which is effected by means of a 
low lift type of an intake valve, may be employed. Otherwise, an 
air-mixing type of a fuel injector may be installed to deliver fuel mixed 
with air. 
The time interval for a rise in exhaust gas temperature is terminated when 
a specified temperature of exhaust gas is reached. The specified 
temperature may be monitored based on the temperature of engine cooling 
water. 
The catalyst is heated to its effective activation temperature by the 
exhaust gas whose temperature is raised by setting the ignition timing of 
the engine at a point after top-dead-center and igniting an air-fuel 
mixture directly after the beginning of an expansion stroke after which 
combustion of the air-fuel mixture initiates. This sequence takes place 
directly after an initial engine start during a time interval for a rise 
in exhaust gas temperature. Consequently, the conversion rate at which 
thermal energy is converted into engine shaft driving power (net power) 
lowers, resulting in an increase in thermal energy loss which leads to a 
rise in exhaust gas temperature. Typically, while retarding the ignition 
timing to a point after top-dead-center would normally result in torque 
fluctuations induced by a reduction in the stability of engine cycling 
caused by lowered combustion efficiency and a drop in net power output, 
nevertheless, according to the exhaust gas purifying system, efficient 
ignition and combustion of an air-fuel mixture is caused, ensuring the 
least stability of cycling stability as needed and keeping torque 
fluctuations to a minimum. 
During the time interval for a rise in exhaust gas temperature, the 
air-fuel ratio is set in a range of 13.5:1 to 18.0:1. As compared to 
conventional engines by which an air-fuel ratio of 12:1 to 13:1 is 
utilized, the amount of latent heat energy needed for fuel atomization and 
the amount of practical heat needed to raise fuel temperature is reduced. 
These reductions in thermal demand are applied to a further rise in 
exhaust gas temperature. In particular, establishing the excessive air 
ratio as more than 1 during the time interval for a rise in exhaust gas 
temperature, that is, providing the air-fuel ratio leaner than the 
stoichiometric air-fuel ratio, decreases the thermal energy needed to 
raise the temperature of fuel and atomize fuel with a resultant rise in 
exhaust gas temperature. Moreover, the amount of hydrocarbon emission 
generated during a cold engine start is decreased. 
Secondary air supplied to the exhaust gas upstream from the catalytic 
converter and heat of the exhaust gas increased by the retardation of the 
ignition timing to a point after top-dead-center aid combustion of 
residual fuel in the exhaust gas. The heat of combustion of the residual 
fuel promotes a further rise in exhaust gas temperature. Moreover, during 
the time interval for a rise in exhaust gas temperature, residual fuel is 
positively burned and the temperature of exhaust gas is increased as a 
result of the virtual exhaust air-fuel ratio being maintained lean at 
higher than a stoichiometric air-fuel ratio of 14.5. Setting the ignition 
timing at a point after top-dead-center is caused after a boost of engine 
speed following a full explosion after engine cranking, making the engine 
operate stably at the setting of the ignition timing at the point after 
top-dead-center. 
Feedback controlling the air-fuel ratio starts during the time interval for 
a rise in exhaust gas temperature, preventing an air-fuel mixture from 
enriching and, consequently, the exhaust gas temperature rise is 
accelerated, while the amount of hydrocarbon emission is reduced. 
During the time interval for the exhaust gas temperature rise after a cold 
engine start, the retardation of ignition timing is adequately determined 
according to the temperature of engine cooling water. Further, during the 
time interval for the exhaust gas temperature rise after a cold engine 
start, the amount of intake air is adequately determined according to the 
temperature of engine cooling water. The amount of intake air is 
increasingly or decreasingly controlled during the time interval for the 
exhaust gas temperature rise to maintains a specified engine speed, 
reducing torque fluctuations even while the temperature of exhaust gas is 
rising during engine warming-up. 
The ignition timing is set on the retarded side after a certain amount of 
time has passed while the temperature of exhaust gas is being raised with 
the effect of decreasing engine speed, resulting in reduced engine noise 
and non-lurching sensation when the vehicle is put into gear. 
The ignition timing is retarded to a point more later but before 
top-dead-center when the vehicle starts under a driving condition while 
the temperature of exhaust gas is being raised than when driving under the 
same driving condition after the exhaust gas temperature rise time 
interval. This provides favorable starting ability and drivability of the 
vehicle. Moreover, while, immediately after the vehicle starts, the engine 
experiences a rapidly increased load and an increased amount of fuel 
consumption, which typically leads to an increased absolute amount of 
hydrocarbon emissions, nevertheless, setting the ignition timing retarded 
by a small angle from an ordinary advanced timing provides a decrease in 
the amount of hydrocarbon emissions. Moreover, heat increased due to an 
increase engine load and the retardation of ignition timing produce the 
effect of raising the temperature of exhaust gas. Even in cases where the 
vehicle starts during the exhaust gas temperature rise time interval, the 
ignition timing is set to a point before top-dead-center when the engine 
is at a high level of output, the vehicle runs with more improved driving 
performance. The ignition timing is, on the other hand, retarded to a 
point after top-dead-center when the engine is at a low level of output, 
the exhaust gas temperature rise is accelerated. 
Generating air turbulence, such as a swirl and a tumble, inside of the 
engine cylinders has the effect of promoting efficient ignition and 
combustion of an air-fuel mixture. A low-lift type of intake valve acts to 
speed the flow of intake air from the intake manifold to the inside of the 
cylinder with the effect of promoting efficient ignition and combustion of 
an air-fuel mixture. Further, the employment of an air mixing type of fuel 
injector which injects a mixture of air and fuel into the intake manifold 
promotes efficient ignition and combustion of the air-fuel mixture. 
Moreover, a high level of air-fuel mixture ignition energy has the purpose 
of promoting efficient ignition and combustion of the air-fuel mixture. 
Simply terminating the time interval for the exhaust gas temperature rise 
after a specified time period has elapsed enables an engine control logic 
and control unit to be simplified. The time interval for the exhaust gas 
temperature rise is terminated after the exhaust gas downstream from the 
catalytic converter has reached a specified temperature, thus raising the 
temperature of exhaust gas until the catalyst attains its effective 
activation temperature and quickly increasing engine net output after the 
activation of the catalyst. The time interval for the exhaust gas 
temperature is terminated when the temperature of engine cooling water, 
which correlates closely with the temperature of exhaust gas. This assures 
a rise in the temperature of exhaust gas until the catalyst certainly 
attains its activation temperature and, after the achievement of 
activation of the catalyst, quickly raises the net output of the engine. 
The ignition timing is set at a point after top-dead center during the time 
interval for the exhaust gas temperature rise after the engine starts in a 
cold condition, igniting an air-fuel mixture first in an expansion stroke. 
Resultingly, the conversion rate of thermal to engine shaft driving energy 
(net power) is lowered, increasing thermal energy loss due to which the 
temperature of exhaust gas rises. Moreover, while setting the ignition 
timing at a point after top-dead-center will generally destabilize engine 
cycling through torque fluctuations induced from worsened combustion 
conditions and lower net power output, the exhaust gas purifying system 
the present invention provides improves ignition and combustion of an 
air-fuel mixture to a point where a minimum level of engine cycling 
stability is maintained and torque fluctuations are adequately controlled. 
Setting the ignition timing at a point after top-dead-center during the 
time interval for the rise in the temperature of exhaust gas after an 
engine cold start causes ignition of an air-fuel mixture for the first 
time in an expansion stroke and thereafter, combustion of the air-fuel 
mixture initiates. Resultingly, the rate of conversion of thermal energy 
to engine shaft driving power (net power) is lowered, resulting in to an 
increase in thermal energy loss increases which always induces a rise in 
the temperature of exhaust gas.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
Referring to the drawings in detail, and in particular FIG. 1, which 
schematically shows overall structure of a reciprocal internal combustion 
engine CE equipped with an exhaust gas purifying system in accordance with 
a preferred embodiment of the present invention, an engine body 1 having a 
plurality of cylinders of which only one cylinder is shown and each of 
which is provided with a combustion chamber 2. A fuel-air mixture is 
supplied through first and second intake ports 3 and 4, and combusted in 
the combustion chamber 2. Burned gases are discharged as exhaust gas 
through first and second exhaust ports 5 and 6 into an exhaust duct 60. 
The first and second intake ports 3 and 4 are opened and closed by first 
and second intake valves 7 and 8, respectively, at a specified timing 
cycle. Similarly, the first and second exhaust ports 5 and 6 are opened 
and closed by first and second exhaust valves 9 and 10, respectively, at a 
specified timing cycle. The air-fuel mixture supplied to the combustion 
chamber 2 is compressed by a piston 11 and ignited by a spark plug 12 (see 
FIG. 3) at a specified timing, namely a crank angle. The time or the crank 
angle at which the ignition plug 12 fires is desirably altered by means of 
an ignition mechanism F which refers to an ignition timing control means 
in this specification. The ignition mechanism F is capable of setting the 
ignition timing in a wide range from a point (MBT) advanced by at least a 
predetermined angle from top-dead-center, where maximum torque is 
generated or the greatest net output power is obtained, to a point 
retarded by a large angle, for example 30.degree. in crank angle, from 
top-dead-center. 
Intake system 13, which is utilized to supply air to each cylinder of the 
engine CE, is comprised of an intake manifold 14 and a common intake duct 
15 which is attached in common to an air inlet of the manifold 14. The 
common intake duct 15 is provided in order from the upstream end an air 
cleaner 16, a hot wire type of air flow sensor 17 and a throttle valve 18. 
An intake air bypass duct 19 is installed in the common intake duct 15 so 
as to allow air to flow bypassing the throttle valve 18. This intake air 
bypass duct 19 is equipped with a relatively small diameter of an idle 
speed control (ISC) duct 20 which adjusts the amount of intake air 
introduced into the engine EC during engine idling, and a relatively large 
diameter of air duct 21 which is used for raising exhaust gas temperature. 
An idle speed control (ISC) valve 22 is installed to the idle speed 
control duct 20 with the purpose of controlling the amount of air entering 
the idle speed control duct 20. Idle speed control is executed by an 
engine control unit (ECU) C as will be discussed later. The exhaust 
temperature rising air duct 21 is equipped with an air control valve 23 
which is duty-controlled by the engine control unit (ECU) C so as to 
adjust the amount of air flowing therein. The bypass air supply system may 
include, in addition to the idle speed control air duct 20 and the exhaust 
temperature rising air duct 21, a third route of air duct (not shown) 
which increases the amount of intake air during cold start of the engine 
or while the engine temperature is low. In this instance, the air duct 21 
must have the largest area of cross section among the three and the third 
air duct must have the smallest area of cross section among the three. 
Accordingly, the idle speed control air duct 20 has an area of cross 
section between those of the others. 
Idling engine speed is maintained at a specified speed by means of 
adjusting the amount of bypass air entering the air duct 20 and the air 
duct 21, respectively, at the air control valve 22 and the air control 
valve 23. These air control valves 22 and 23 are controlled by means of 
the engine control unit C, thus allowing idling speed to be desirably set 
by the engine control unit C. The bypass air supply system consisting of 
these air ducts 19, 20 and 21 and air control valves 22 and 23 are 
referred to in this specification as an intake air volume control means. 
A fuel injector 24 is installed to the first intake port 3 with the purpose 
of injecting fuel into the air in the first intake port 3. A turbulence 
control valve 25 is installed to the second intake port 4 with the purpose 
of opening and closing the second intake port 4. This turbulence control 
valve 25 is operated by means of a vacuum or negative pressure response 
actuator 26. A solenoid valve 29, disposed between a vacuum tank 28 which 
accumulates vacuum pressure supplied from a surge tank 27 installed within 
the intake manifold 14 and the actuator 26, controls supply and discharge 
of the vacuum pressure from the vacuum tank 28 to operate the actuator 26. 
Ignition timing (which is represented by a crank angle) of fuel injection 
and the amount of fuel injected by the fuel injector 24 are controlled by 
means of the engine control unit (ECU) C in response to operating 
conditions of the engine CE. Engine control unit (ECU) C sets the amount 
of fuel injection through the injector 24 based on the amount of intake 
air entering the engine CE which is monitored by the air flow sensor 17, 
thereby allowing the air-fuel ratio of a fuel mixture supplied to the 
combustion chamber 2 to be desirably adjusted. For the sake of 
convenience, the term "in-cylinder air-fuel ratio" as used herein shall 
mean and refer to the air-fuel ratio at which a fuel mixture is supplied 
to the combustion chamber, and the term "secondary air supplied air-fuel 
ratio" or "exhaust air-fuel ratio" as used herein shall mean and refer to 
the virtual air-fuel ratio which is the ratio of the sum of intake air, 
including the amount of intake air supplied to the combustion chamber and 
the amount of intake air added as secondary air to the exhaust, relative 
to the amount of injected fuel. 
Fuel injector 24 is of an air mixture-type which forms fine fuel particles 
in the proximity of the fuel discharge nozzle by supplying mixing air. A 
fuel is supplied by a fuel pump 31 to the fuel injector 24 through a fuel 
supply line 33 from a fuel tank 30 via a fuel filter 32 and returned to 
the fuel tank 30 through a fuel return line 35 via a pressure regulator 
valve 34. The fuel injector 24 cooperates with a mixing air supply system 
and a fuel vapor supply system which are well known in various types in 
the art and take any known form. Although not shown in detail in the 
figure, in the air-mixing type fuel supply system, the mixing air supply 
system for the fuel injector 24 disposed at the first intake port 3 of 
each cylinder includes a mixing air duct 37 equipped with an air control 
valve 36. The mixing air duct 37 is connected at its upstream end to the 
intake bypass duct 19 and at its branched downstream ends to the 
respective fuel injectors 24, is installed on a branch at the air outlet 
side of supply duct 37 at each cylinder. Air control valve 36 is opened 
and closed by the engine control unit C in accordance with operating 
conditions. For example, when the engine CE is completely warmed up, the 
air control valve 36 is closed during idling or when the engine CE is 
running under high engine loads and is otherwise open to supply mixing air 
to the fuel injector 24. 
Fuel vapor supply system includes a canister 41 which collects or traps 
fuel vapors discharged from the fuel tank 30 and a purge duct 42, provided 
with a purge solenoid valve 43, which extends from the canister 41 and is 
connected to a mixing chamber 40 formed upstream from a point at which the 
mixing air duct 37. The fuel vapors collected or trapped in the canister 
41 enter the purge duct 42 when the purge solenoid valve 43 activates and 
are supplied to the injector 24 in close proximity to the injection 
nozzle, thus allowing the fuel vapors to enter the combustion chamber 2 
during intake strokes. 
Purge solenoid valve 43, which is activated by the engine control unit C 
when predetermined purge conditions are met, allows an amount of purged 
fuel vapors specified according to engine operating conditions to enter 
the air-mixing system. The purge solenoid valve 43 is operated at, for 
instance, a duty rate of 20% when the amount of purged fuel vapors is 
established to be constant in a specified range of engine operating 
conditions where a stoichiometric air fuel mixture is maintained in spite 
of engine speeds and engine loads. On the other hand, the purge solenoid 
valve 43 is controlled in response to the amount of intake air detected by 
the air flow sensor 17 and engine speed so that the amount of purge fuel 
vapors is reduced during a lean burn driving condition where the air-fuel 
ratio is on the lean side as compared with during driving conditions where 
the stoichiometric air fuel mixture is set range and that the rate of 
reduction in the amount of purged fuel vapors becomes larger with a 
decrease in the amount of intake air. In this manner, air-mixing fuel 
injection by the injector 24 acts to break fuel into fine particles, 
providing a uniform mixture of air and fuel that can be ignited and 
combusted with greater efficiency while reducing nitrogen oxides (NOx) 
emissions during lean burn driving conditions. 
While not shown in the figure, the ignition mechanism F, which supplies 
electrical power to the spark plug 12 to ignite the air-fuel mixture, can 
be designed as a high-energy type of ignition system which produces a high 
energy spark. As a practical example, the high-energy type ignition system 
is designed and adapted so as to provide increased ignition energy as a 
result of the characteristics of ignition coil altered as compared to 
those for ordinary engines and to extend a spark duration 1.5 times as 
long as that of ordinary ignition systems by making use of low-resistance 
spark plug cables which reduce transmission losses. Together, the spark 
plug 12 is of platinum construction which is resistant to high voltages. 
The ignition mechanism F thus constructed provides a great improvement of 
the efficiency of ignition and combustibility. 
Exhaust system of the engine CE is equipped with a linear oxygen (O.sub.2) 
sensor 44 which monitors the concentration of oxygen (O.sub.2) in the 
exhaust gas, a catalytic converter 45 for controlling exhaust emissions, 
an air duct 61 and a lead valve 62. The air duct 61 supplies secondary 
air, of which the amount is controlled by the lead valve 62, to the 
exhaust gas in the exhaust duct 60 upstream from the catalytic converter 
45. In this instance, the engine control unit C uses the output signal 
from the linear oxygen sensor 44 to compute an air-fuel ratio based on the 
concentration of oxygen (O.sub.2) in the exhaust gas represented by the 
output signal. The engine control unit C also operates the lead valve 62 
to control the amount of secondary air to be supplied to the exhaust 
system. While the secondary air is supplied to the exhaust system with the 
purpose of burning residual unburned fuel in the exhaust gas, the 
secondary air supplied to the exhaust duct 60 aids in the combustion of 
unburned fuel in the exhaust gas accompanied by an additional result of 
raising the temperature of the exhaust gas. 
The exhaust gas purifying catalyst element employed in the catalytic 
converter 45 is of a three-way type incorporating a silicon oxide carrier 
containing precious metals such as platinum, rhodium and palladium. This 
type of exhaust gas purifying catalyst element provides adequate exhaust 
gas purifying performance at temperatures only above a specified 
activation temperature of, for example, 350.degree. C. 
Alternatively, what is called a "new three-way catalyst" may be employed as 
the exhaust gas purifying catalyst element. The new three-way catalyst 
provides the capability to reduce nitrogen oxides (NOx) at air-fuel 
mixtures leaner than the stoichiometric air fuel mixture. As an example, 
the catalyst carrier of, desirably, zeolite incorporates precious metals 
such as iridium and platinum, or iridium, platinum and rhodium, as 
activation agents. The new three-way catalyst provides adequate emission 
control of nitrogen oxides (NOx) and purification of the exhaust gas even 
at lean air-fuel ratios. Specifically, in many instances where ordinary 
three-way catalysts are employed, under the stoichiometric air-fuel ratio, 
the exhaust gas contains only oxygen (O.sub.2) too low in level to oxidize 
hydrocarbons (HC) and carbon monoxide (CO), and as a result, the oxygen 
(O.sub.2) in the nitrogen oxides (NOx) is consumed in the reaction and the 
nitrogen oxides (NOx) is deoxidized or broken down. On the other hand, 
under lean air-fuel ratios, the relatively large amount of oxygen 
(O.sub.2) in the exhaust gas reacting with the hydrocarbons (HC) and 
carbon monoxide (CO) results in the nitrogen oxides (NOx) being left 
unreacted, thus causing the nitrogen oxides (NOx) to be discharged without 
being deoxidized. 
Contradistinctively, the new three-way catalyst utilizing a zeolite carrier 
which is highly porous results in capturing a large amount of hydrocarbons 
(HC) in pores of the zeolite carrier, while the nitrogen oxides (NOx) 
concurrently adhere to the precious metals incorporated by the zeolite 
carrier. Even with a large amount of oxygen (O.sub.2) in the exhaust gas, 
the hydrocarbons (HC) captured by the zeolite carrier react with the 
oxygen (O.sub.2) component of the nitrogen oxides (NOx) in the catalytic 
reaction as a result of the hydrocarbons (HC) existing in close proximity 
to the nitrogen oxides (NOx), resulting in that the catalyst eliminates 
the nitrogen oxides (NOx) with a high efficiency even under lean air-fuel 
ratios. In this case, iridium is utilized as a means of refining the 
active precious metals, thus increasing the exhaust gas purification 
efficiency of the catalyst and extending the service life of the catalyst. 
Exhaust system, which is provided with the catalytic converter 45 of an 
under foot type in the above embodiment and in which secondary air is 
supplied to the catalytic converter 45, may be modified as shown in FIG. 
28. 
Referring to FIG. 28, a pre-catalytic converter 80 may be installed to the 
exhaust duct 60 in close proximity to the engine CE, in other words, 
attached directly to an exhaust manifold. Further, in place of the linear 
oxygen (O.sub.2) sensor, a lambda oxygen (.lambda.O.sub.2) sensor 81 of a 
self-heating type having an incorporated heater is installed to the 
exhaust duct 60 before or upstream from the pre-catalytic converter 80. In 
the exhaust system, secondary air is not supplied to the exhaust gas. The 
same three-way catalyst as used in the catalytic converter 45 may be 
employed in the pre-catalytic converter 80. As is well known in the art, 
the lambda oxygen (.lambda.O.sub.2) sensor 81 does not monitor a linear 
change in air-fuel ratio but detect whether the air-fuel ratio is leaner 
or richer than the stoichiometric air-fuel ratio. 
Since the exhaust gas does not change in temperature almost at all until 
entering the pre-catalytic converter 80, the modified exhaust system 
enables the catalyst to take a time to become active shorter as compared 
with the exhaust system not being provided with a pre-catalytic converter 
and, consequently, yields an improved efficiency of purifying the exhaust 
gas. As a result, as will be described later, the retardation of ignition 
timing during a time interval for which the temperature of exhaust gas is 
enforced to rise can be decreased, thereby improving the stability of 
engine operation. The self-heating type of lambda oxygen (.lambda.O.sub.2) 
sensor 81 quickly becomes sensitive to oxygen in the accelerated exhaust 
temperature raising period. It is judged that the lambda oxygen 
(.lambda.O.sub.2) sensor 81 rightly becomes active upon an occurrence of a 
change in the sensor output from a level indicative of a lean air-fuel 
ratio to a level indicative of a rich air-fuel ratio. The fuel mixture is 
prevented from being overly enriched during the accelerated exhaust 
temperature raising period by performing open-looped air-fuel ratio 
control at the beginning of the accelerated exhaust temperature raising 
period and subsequently air-fuel ratio feedback control after the 
activation of the lambda oxygen (.lambda.O.sub.2) sensor 81. In this 
manner, a raise in exhaust temperature is more enhanced, the discharged 
amount of hydrocarbons (HC) is reduced more during the accelerated exhaust 
temperature raising period. 
In order to monitoring the operating condition of the engine CE, the engine 
CE is equipped, in addition to the air flow sensor 17 and the linear 
oxygen (O.sub.2) sensor 44, with a throttle position sensor 46 which 
monitors open positions of the throttle valve 18, an idle switch 47 which 
detects the throttle valve 18 is in the closed position, temperature 
sensors 48 and 49 which monitor the temperatures of intake air and engine 
cooling water, respectively, a knock sensor 50 capable of monitoring an 
occurrence of knocking of the engine CE, and a temperature sensor 63 which 
monitors the temperature of exhaust. Further, a distributor 51, which 
forms a part of the ignition system, is equipped with an angle sensor 52 
which is capable of monitoring a turned angle of a crankshaft, and a 
cylinder discrimination sensor 53 which is capable of outputting a signal 
based on which the cylinders are differentiated. 
Output signals from all of the sensors and switches are sent to the engine 
control unit (ECU) C equipped with a microcomputer. The engine control 
unit (ECU) C processes the information contained in the output signals and 
controls operations of the ignition system F, the idle speed control (ISC) 
valve 22, the air control valve 23, the fuel injector 24, the turbulence 
control valve 25, the solenoid valve 29, the air control valve 36, and the 
purge solenoid valve 43. Specifically, the engine control unit C sets an 
ignition timing through the ignition system F and controls engine speed 
during idling by regulating the amount of intake air by means of the idle 
speed control (ISC) valve 22 and air control valve 23, controls an 
air-fuel ratio or the amount of fuel and a fuel injection timing through 
the fuel injector 24, controls the strength of swirl and tumble 
(ignitability and combustibility) of the fuel mixture in the combustion 
chamber 2 by regulating opening of the turbulence control valve 25, 
controls supply of the mixing air to the fuel injector 24 through the air 
control valve 36, and controls supply of the secondary air through the 
lead valve 62. The following explanation makes reference to FIGS. 2 
through 5. Turbulence is generated in the combustion chamber 2 to improve 
the ignition and combustion characteristics of an air-fuel mixture. Swirl 
(horizontal vortex turbulence) and tumble (vertical vortex turbulence) are 
utilized to speed the flow of an air-fuel mixture into the combustion 
chamber 2 with the purpose of improving ignition and combustion 
characteristics. As will be further explained later, because the ignition 
and combustion characteristics of an air-fuel mixture are improved in the 
manner stated above, the stability of engine cycling is obtained and 
torque fluctuations are kept to a minimum, even with an ignition timing 
retarded greatly to a point after top-dead-center during cold starting of 
the engine CE. 
As shown in FIGS. 2 through 5, each cylinder of the engine (only one 
cylinder is shown in the figures) is equipped with the fuel injector 24 in 
close proximity to the first intake port 3, the amount of injected fuel 
and timing of fuel injection of the fuel injector 24 being determined by 
the engine control unit C. Turbulence control valve 25 is installed to the 
second intake port 4, the opening of the turbulence control valve 25 being 
determined by the engine control unit C in response to engine operating 
conditions. Valve seats 55 and 56 are formed in openings 3a and 4a to the 
combustion chamber 2 at the ends of the first intake port 3 and second 
intake port 4, respectively. Similarly, valve seats 57 and 58 are formed 
in openings 5a and 6a to the combustion chamber 2 at the ends of the first 
exhaust port 5 and second exhaust port 6, respectively. The first and 
second intake valves 7 and 8 are installed to the first and second intake 
ports 3 and 4, respectively, formed on a ceiling 59 (see FIGS. 4 and 5) of 
the combustion chamber 2. Similarly, the first and second exhaust valve 9 
and 10 are installed to the first and second exhaust ports 5 and 6, 
respectively, formed on the ceiling 59 of the combustion chamber 2. The 
openings 3a and 4a at the ends of the first and second intake ports 3 and 
4 are arranged in a left half section of the combustion chamber 2 as 
viewed in a cross section, and the openings 5a and 6a at the ends of the 
first and second exhaust ports 5 and 6 are likewise arranged in a right 
half section of the combustion chamber 2 as viewed in the same cross 
section. The spark plug 12 is installed in the ceiling 59 of the 
combustion chamber 2 at an off-center position adjacent to the exhaust 
ports 5 and 6. As will be explained later, the first intake port 3 is 
configured as a swirl port which provides a swirl of fuel mixture having a 
swirl ratio greater than 1.0 to the combustion chamber 2, and the second 
intake port 4 is configured as a tumble port which provides a tumble of 
fuel mixture having a tumble ratio greater than 1.5 to the combustion 
chamber 2. The first intake valve 7 is of a low-lift type which have a 
valve lift set to be relatively small, so that the speed of an air-fuel 
mixture traveling through a space between the first intake port 3 and 
first intake valve 7 to combustion chamber 2 is increased and, as a 
result, the turbulence generating effect in the combustion chamber 2 is 
further promulgated. This provides a great improvement of the efficiency 
of ignition and combustibility. The second intake valve 8 may be of the 
low-lift type. 
As is apparent from FIG. 2, the first intake port 3 has approximately 
straight part 3b incorporated at its upstream side thereof and slightly 
curved part 3c incorporated at its downstream side thereof. In the same 
manner, the second intake port 4 has approximately straight part 4b 
incorporated at its upstream side thereof and slightly curved part 4c 
incorporated at its downstream side thereof. Straight part 3b of the first 
intake port 3 establishes a specified angle .alpha. (which is hereafter 
referred to as a first intake port inclination angle) between an axis 
L.sub.1 of the straight port 3b and a cross sectional plane A.sub.1 
perpendicular to the cylinder center axis L.sub.2. Similarly, the straight 
part 4b of the second intake port 4 establishes a specified angle .alpha.' 
(which is hereafter referred to as a second intake port inclination angle) 
between an axis L.sub.3 and the cross sectional plane A.sub.1. First 
intake valve 7, comprised of a valve stem 7a and a valve head 7b, forms an 
angle .theta. (which is hereafter referred to as an intake valve 
inclination angle) between an axis L.sub.4 of the valve stem 7a and the 
cylinder center axis L.sub.2. The valve head 7b has the under surface 
oriented parallel with the pent roof ceiling 59 of the combustion chamber 
2 and incorporates a specified valve face angle .beta. between its upper 
and lower surfaces. Similarly, the first exhaust valve 9, comprised of a 
valve stem 9a and a valve head 9b, form an angle .theta.' (which is 
hereafter referred to as an exhaust valve inclination angle) between an 
axis L.sub.5 of the valve stem 9a and the cylinder center axis L.sub.2, 
The valve head 9b has the under surface oriented parallel with the pent 
roof ceiling 59 of the combustion chamber 2. In the following description, 
the term "port incident angle .gamma." shall mean and refer to an angle 
formed between the axes L.sub.1 and L.sub.4, and the term "port angle 
difference .omega." is defined by a deference between the first and second 
intake port inclination angles (.alpha.-.alpha.'). 
As apparently understood from FIGS. 4 and 5, the combustion chamber 2 is 
formed with the pent roof type of ceiling 59 (depicted by contour lines) 
having an angle .theta." relative to the cylinder horizontal cross section 
A.sub.1 on the exhaust side. In this instance, the under surface of the 
valve head 9b of first exhaust valve 9 also forms an angle .theta." 
relative to the cylinder horizontal cross section A.sub.1. Accordingly, 
the angles .theta." and .theta.' are equivalent. The pent roof type of 
combustion chamber 2 has the ceiling angle .theta." relative to the 
cylinder horizontal cross section A.sub.1 on the exhaust side which is 
relatively small. Together, the pent roof type of ceiling 59 has also a 
relatively small angle relative to the cylinder horizontal cross section 
A.sub.1 on the intake side. Accordingly, the first and second intake 
valves 7 and 8 and the first and second exhaust valves 9 and 10 have 
center axes of their valve stems oriented on more a vertical axis. 
The angles of .alpha., .theta., .theta.'(.theta.") and .beta. are desirably 
established so as to satisfy all of the following inequalities. 
EQU .beta.&gt;.theta. (I) 
EQU .alpha.&lt;.theta.' (II) 
EQU .alpha..ltoreq..theta. (III) 
The parameters set forth are necessary for suitably producing a swirl in 
the combustion chamber 2. The first parameter (.beta.&gt;.theta.) defines the 
direction of an air-fuel mixture flow. If the valve face angle .beta. is 
less than the specified angle .theta., the direction of the air-fuel 
mixture flow into the combustion chamber 2 from the first intake port 3 is 
deflected by the upper surface of the valve head 7b and resultingly hits 
the ceiling 59 of the combustion chamber 2, so that the inflow speed of 
the air-fuel mixture into the combustion chamber 2 is lowered, thereby 
preventing the swirl effect from forming. It is, however, not desirable to 
increase the valve face angle .beta. to a great extent as such an increase 
would make the cross sectional area between the first intake port 3 and 
valve head 7b extremely narrow, thus increasing inflow friction. The 
second and third parameters (.alpha.&lt;.theta.' and .alpha..ltoreq..theta.) 
yields the distribution of inflow speed components of the air-fuel mixture 
into the combustion chamber 2. If the first intake port inclination angle 
.alpha. is made as small as possible, the horizontal inflow speed 
component (the component of inflow speed in the direction parallel to the 
cylinder cross section A.sub.1) of the air-fuel mixture entering the 
combustion chamber 2 from the first intake port 3 is increased with an 
effect of accelerating an occurrence of a swirl. Satisfaction of these 
parameters ensures the air-fuel mixture to flow into the combustion 
chamber 2 from the first intake port 3 without interfering with the 
combustion chamber ceiling 59 when the piston is in its down stroke and 
adequately maintain the horizontal speed, so as to yield an accelerated 
occurrence of a swirl in the combustion chamber 2. 
As a result of the configuration, a strong swirl of a swirl ratio SR of 
more than 1.0 is formed in the combustion chamber 2 when the turbulence 
control valve 25 is closed or constricted (partially open), thus 
significantly improving the ignition and combustion process of the 
air-fuel mixture. The swirl ratio SR is generally defined as the number of 
horizontal swirl circulations of an air-fuel mixture in the cylinder 
divided by the number of revolutions of the engine. The number of 
horizontal swirl circulations of an air-fuel mixture is determined, as 
shown in FIG. 6 for example, by means of an impulse swirl meter 65 
installed at a position F.sub.2 at a distance of 1.75 D from the lower 
surface of a cylinder head F.sub.1 for an engine having a cylinder bore 
diameter of D. The torque driving the impulse swirl meter 65 (which is 
referred to as the impulse swirl meter torque G) is used to compute the 
swirl rate in a manner well known in the art. In FIG. 6, denoted by a 
reference F.sub.3 is the top of piston 11 at bottom-dead-center. 
Impulse swirl meter torque is measured in the following manner. That is, 
the swirl energy acting on the piston top is regenerated by the impulse 
swirl meter 65 installed at the position F.sub.2 at the distance of 1.75 D 
from the lower surface of the cylinder head F.sub.1 so as to uncover how 
strong the swirl circulating energy is around the piston top. The impulse 
swirl meter 65 incorporates a multi-cell honeycomb construction to which 
the swirl directional force applied to each cell is used to compute the 
total impulse swirl meter torque G. 
To explain in more detail, assuming the air-fuel mixture is flowing into 
the combustion chamber 2 during a period until the piston 11 moves to 
bottom-dead-center after the intake valve opens; the circulation speed of 
the air-fuel mixture circulating along the periphery of the combustion 
chamber 2 becomes the greatest speed when the piston reaches its 
bottom-dead-center. Accordingly, the swirl ratio SR is obtained by 
computing the total amount of angular motion at each specified crank angle 
from the time the intake valve opens until the piston reaches its 
bottom-dead-center. Based on this principle, the swirl ratio SR is 
computed from the following formulas (IV) and (V). 
EQU SR=.eta.v.multidot.D.multidot.S.multidot..PI.(cf.multidot.Nr.multidot.d.mu 
ltidot..phi.)!/n.multidot.d.sup.2 
.multidot.(.PI.cf.multidot.d.multidot..phi.).sup.2 ! (IV) 
EQU Nr=8.multidot.G/(M.multidot.D.multidot.V.sub.0) (V) 
In the above formulas: 
SR is the swirl ratio; 
.eta.v is the volumetric efficiency (=1); 
D is the diameter of the cylinder bore; 
S is the stroke 
n is the number of the intake valves 
d is the diameter of the throat 
cf is the flow volume coefficient in relation to the valve lift; 
Nr is the dimensionless rig swirl value in relation to the valve lift; 
.phi. is the crank angle; 
G is the impulse swirl meter torque; and 
V.sub.0 is the speed head. 
The formula (V) is induced in the following order: 
EQU G=I.multidot..omega.r (1) 
EQU I=M.multidot.D.sup.2 /8 (2) 
Substituting formula (2) into formula (1): 
EQU G=M.multidot.D.sup.2 .multidot..omega.r/8 (3) 
Rearranging formula (3): 
EQU D.multidot..omega.r=8.multidot.G/(M.multidot.D) (4) 
The dimensionless rig swirl value is expressed by the following formula: 
EQU Nr=D.multidot..omega.r/V.sub.0 (5) 
When formula (4) is factored into formula (5), the formula (V) is given as 
follows: 
EQU Nr=8.multidot.G/(M.multidot.D.multidot.V.sub.0) (V) 
In the above formulas, .omega.r is the rig swirl value. 
In regard to the second intake port 4, the second intake port inclination 
angle .alpha.' is established to be a relatively large value. Because the 
air-fuel mixture entering the combustion chamber 2 from the second intake 
port 4, when the turbulence control valve 25 is open, has a relatively 
large component of speed at which it is forced downward, a strong tumble 
(vertical vortex) having a tumble ratio RT of more than 1.5 is generated 
in the downward moving air-fuel mixture. While the tumble creates promoted 
turbulence of the air-fuel mixture in the combustion chamber 2 and 
improves efficient ignition and combustion of the air-fuel mixture, the 
turbulence created in the combustion chamber 2 also suppresses an 
occurrence of knocking. For example, while knocking would usually be apt 
to occur in a range of engine operations under moderate or high engine 
loads where the turbulence control valve 25 is open, the tumble effect 
explained above prevents the occurrence of knocking. 
Tumble ratio RT is generally defined as the number of vertical vortex 
circulations of an air-fuel mixture in the cylinder divided by the number 
of revolutions of the engine. The tumble ratio RT is computed in the 
substantially same manner as explained previously in regard to the swirl 
ratio RS. 
In reference to FIG. 3 which is a plan view of the top of the combustion 
chamber 2, the opening 3a of first intake port 3 is detected toward or 
faces the inner periphery of combustion chamber 2, and the opening 4a of 
second intake port 4 is detected toward nearly the center of combustion 
chamber 2. Resultingly, a swirl turbulence effect is enhanced by the first 
intake port 3, and a tumble turbulence effect is enhanced by the second 
intake port 4. On the other hand, the construction of these intake ports 3 
and 4 prevents a reverse swirl effect from occurring at the second intake 
port 4. In this manner, in the combustion chamber 2, a swirl having a 
swirl ratio of more than 1.0 and a tumble having a tumble ratio of more 
than 1.5 are produced such as shown in FIG. 5 by arrows X.sub.1 and 
X.sub.2, respectively, thereby significantly improving the ignition and 
combustion characteristics of an air-fuel mixture. 
Engine control unit (ECU) C performs catalyst activation control during 
cold start of the engine CE in which the catalytic converter 45 becomes 
operative with an expected exhaust gas purifying efficiency or emission 
control performance as quickly as possible by means of promoting a raise 
in the temperature of the catalytic element of catalytic converter 45 to 
the effective active temperature of catalyst. 
In the catalyst activation control, the engine control unit (ECU) C 
controls the ignition mechanism F including the spark plug 12, the air 
control valve 23, the fuel injector 24, the turbulence control valve 25, 
the solenoid valve 29, and the air control valve 36 based on the 
information contained in the output signals of the sensors and switches 
previously described, respectively. As specifically stating, the engine 
control unit (ECU) C sets the ignition timing of spark plug 13 through the 
ignition mechanism F, controls the speed of rotation of the engine CE by 
means of regulating the amount of bypass air through the idle speed 
control (ISC) valve 22 and air control valve 23, controls the amount of 
fuel injected through the injector 24, controls the strength of swirl and 
tumble in the combustion chamber 2 by opening and closing the turbulence 
control valve 25, controls supply of mixing air to the injector 24 through 
the operation of air control valve 36, and controls the amount of 
secondary air supplied to the exhaust duct 60 through the lead valve 62. 
The catalyst activation control is executed according the following basic 
aspects. In the catalyst activation control during cold starting where the 
engine CE is still at ambient temperatures, while ensuring the stability 
of engine cycling and keeping torque fluctuations to a minimum, the 
temperature of the catalytic element for exhaust purification is quickly 
raised to the effective active temperature of, for example, above 
350.degree. C. within a time, for example 20 seconds, from the engine 
start, by means of adequately controlling various control factors such as 
the ignition timing, the air-fuel ratio, the speed of engine or the amount 
of bypass air, the amount of secondary air, the strength of turbulence, 
i.e. swirl and tumble, in the combustion chamber, the amount of mixing 
air, and the energy of ignition spark. By virtue of the catalyst 
activation control, the effective active temperature is reached quickly 
without the aid of a heater for raising the temperature of the catalytic 
element, such as an afterburner, an EHC, or other like incorporated in the 
exhaust system. In this instance, the determination of a cold start of the 
engine CE, or the determination as to whether the engine CE is still at 
ambient temperatures or not, is undertaken by means of the water 
temperature sensor 49 which detects whether the temperature of engine 
cooling water is below or above the specified temperature. The factors 
influential in the temperature of exhaust gas during a cold start 
condition basically include the ignition timing, the air-fuel ratio, the 
speed of engine or the amount of bypass air, and the amount of secondary 
air. In other words, the temperature of exhaust gas is raised by retarding 
the ignition timing (see FIGS. 7, 8 and 13), setting the air-fuel ratio 
closely to the stoichiometric air-fuel ratio (see FIG. 13), raising the 
speed of engine or the amount of bypass air (see FIG. 8), and supplying 
secondary air to the exhaust duct 60 upstream from the catalytic converter 
45 (see FIG. 14). However, retarding the ignition timing to a point well 
after top-dead-center will adversely affect the ignition and combustion 
efficiency and induce torque fluctuations which result from the lowered 
stability of engine cycling. It is for this reason that the exhaust gas 
purifying system according to the aspects above described provides a 
significant increase in ignition and combustion efficiency so as to 
effectively suppress torque fluctuations through the ensured stability of 
engine speed with the aid of increasing the swirl ratio to more than 1.0, 
increasing the tumble ratio to more than 1.5, increasing the speed of 
intake air flowing into the combustion chamber 2, supplying assist air to 
the fuel injector 24, and raising the electrical energy supplied to the 
spark plug 12. 
One of the motives for quickly attaining the activation of the catalyst 
after a cold start of engine lies in the tightened automotive emission 
regulations or standards in, for instance, the United States and other 
countries. In the near future, the United States emission regulations or 
standards will demand 1/6 the current allowable emission of hydrocarbons 
(HC) and 1/2 the current allowable emission of nitrogen oxides (NOx). In 
order to meet these standards, it becomes necessary to achieve the 
complete activation of the catalyst before the vehicle is driven while the 
engine is not yet warmed up. In other words, hydrocarbon emissions will 
increases significantly if the vehicle is driven before the active 
temperature is attained by the catalyst. Together, according to the 
emissions test mode in the United States, at least a 20-second warm-up 
time is called for before the vehicle is initially driven after a cold 
engine start. Therefore, in light of the emissions standards in the United 
States, it becomes necessary for the catalyst to attain and exceed the 
effective active temperature within 20 seconds after an engine start. 
The catalyst activation control significantly retards the ignition timing 
by a specified angle of, for example, 30.degree. in crank angle after 
top-dead-center for a specified period of time necessary to raise the 
temperature of exhaust gas during a cold engine start so as to 
intentionally lower the conversion rate of the thermal energy from 
combustion of the air-fuel mixture to net power output and utilize the 
resultant heat loss to raise the temperature of the exhaust gas, thereby 
promoting a fast rise in the temperature of the catalyst to or above the 
effective activation temperature. In other words, the engine CE itself is 
made to function much like an afterburner to raise the temperature of 
exhaust gas. 
It becomes possible to quickly raise the temperature of exhaust gas to or 
above the effective activation temperature by significantly retarding the 
ignition timing by a specified angle in crank angle after top-dead-center 
so that a range where an increasing rate of in-cylinder pressure for each 
specified crank angle is more than 0 (zero) is established after the 
middle of an expansion stroke. 
FIGS. 15 and 16 show characteristics of in-cylinder pressure change in 
relation to crank angles with use of ignition timing as a parameter for 
air-fuel ratios of 16.7 and 13.0, respectively. In FIGS. 15 and 16, the 
lines showing the parameters of 10, 0, -5, -10, -15, -20 and -23 denote 
various degrees of timing advance in crank angle, while the broken line 
shows in-cylinder pressure without combustion of an air-fuel mixture in 
the combustion chamber 2. As is apparent from FIGS. 15 and 16, if the 
ignition timing is retarded at a point after top-dead-center, the range 
where an increasing rate of in-cylinder pressure for each specified crank 
angle is more than 0 (zero) or a range in which a second peak of 
in-cylinder pressure lies is established after the middle of an expansion 
stroke. 
In this manner, in cases where the ignition timing is set to a point after 
top-dead-center, particularly in cases where the ignition timing is 
retarded by a crank angle of -10.degree., the combustion of an air-fuel 
mixture reaches its peak after a significant drop in in-cylinder pressure, 
thus causing the thermal energy to be converted into net output power at a 
significantly low rate. As a result, this considerable increase in thermal 
energy loss takes the form of a significant increase in exhaust gas 
temperature. 
The reasons for the significant rise in exhaust gas temperature resulting 
from the retardation of ignition timing to a point after top-dead-center 
are cited below. 
As a general rule, a large proportion, for instance approximately 70 
percent, of the thermal energy generated from combustion of an air-fuel 
mixture in the combustion chamber 2 is given up to thermal energy loss in 
various forms such as exhaust loss which acts as a heat emission to a low 
heat source necessary to effect the thermal cycle, heat loses from a 
cooling effect which protects structural material of the engine from heat 
damage, mechanical friction loss, and other types of thermal loss. The 
remaining thermal energy is converted to net power output. Therefore, it 
is natural to assume that the temperature of exhaust gas will rise as 
thermal loss to the exhaust gas increases. Together, in this heat exchange 
or accumulation and expenditure, the cooling loss and mechanical friction 
loss do not change so significantly according to engine operating 
conditions. Consequently, there is a correlation between thermal loss to 
the exhaust and net power output that an increase in either one will 
result in a complementary decrease in the other. 
As a general rule, the output torque, i.e. net power output, of an engine 
changes as the ignition timing changes as shown in FIG. 7. Maximum net 
power output is attained only at a specified ignition timing (MBT) 
advanced by a specific crank angle of, for instance, 20.degree. before 
top-dead-center. In a conventional engine, the ignition timing is retarded 
as compared with the advanced ignition timing (MBT) in the present 
invention, for instance at a crank angle of 10.degree. before 
top-dead-center, during a cold engine start. In this instance, advancing 
or retarding the ignition timing from the advanced ignition timing (MBT) 
results in lowered net power output, the net power output being more 
significantly reduced in particular when the ignition timing is retarded, 
and the reduction in net power output complementarily increases thermal 
loss accompanied by an increase in exhaust gas temperature. FIG. 7 shows, 
in addition to change in engine output torque, one example of changes in 
exhaust gas temperature in relation to ignition timing. 
FIG. 8 shows one example of characteristics of change in exhaust gas 
temperature in relation to ignition timing for various engine speeds. FIG. 
9 shows characteristic curves of change in estimated catalyst temperature, 
20 seconds after an initial engine start, in relating to ignition timing 
for various engine speeds. FIGS. 8 and 9 prove the pronounced effect of 
ignition timing on exhaust and catalyst temperature. As shown in FIG. 8, 
there is a significant sharp rise in exhaust gas temperature when the 
ignition timing is set to a point after top-dead-center. During a cold 
start of the engine CE, the heat generated by combustion of the air-fuel 
mixture is lost to the cold surfaces of the combustion chamber 2, 
especially to the ceiling of the combustion chamber comprised of the under 
surface of the cylinder head. On the other hand, if retarding the ignition 
timing to a point significantly after top-dead-center, there occurs the 
effect of shortening the duration of thermal exchange with surfaces of the 
combustion chamber 2, specifically the ceiling of the combustion chamber 2 
which occupies the greater part of the combustion chamber surfaces, so as 
to lower the heat exchange rate, or the rate of cooling loss to the engine 
CE with a result of increasing the rate of thermal loss to the exhaust 
gas, thereby raising the temperature of the exhaust gas. 
According to the aspects of the present invention described above, when the 
engine CE starts in a cold condition under a light load, while maintaining 
only the minimal level of net power output necessary for the engine CE to 
operate itself, the ignition timing is variably retarded in a range after 
top-dead-center as late as possible during a specific period of time for 
which acceleration of a rise in the temperature of exhaust gas is made. 
This retardation of the ignition timing causes great exhaust loss with the 
effect of increasing a rise in the temperature of exhaust gas and thus 
quickly raising the temperature of the exhaust purifying catalyst to a 
point above the effective activation temperature. However, significantly 
retarding the ignition timing causes aggravation of the ignition and 
combustion efficiency or characteristics of an air-fuel mixture with a 
lack of net power output, resulting in a decrease in engine cycling 
stability which always causes torque fluctuations. In these circumstances, 
the exhaust gas purifying system of the present invention is designed and 
adapted to promote shaping and burning an air-fuel mixture when the 
ignition timing is retarded in a range significantly after top-dead-center 
during the period of time for which acceleration of a rise in the 
temperature of exhaust gas is made, thereby ensuring the stability of 
engine cycling and suppressing an occurrence of torque fluctuations. 
Practically, a swirl having a swirl ratio of more than 1.0 is created in 
the combustion chamber 2 so as to increase ignition and combustion 
efficiency of an air-fuel mixture. In regard to obtaining a strong swirl 
of a swirl ratio above 1.0, closing or partially closing the second intake 
port 4 by the turbulence control valve 25 will result in the desirable 
condition of the air-fuel mixture entering the combustion chamber 2 
completely or mainly through the first intake port 3 which functions as 
the swirl port as described previously. In this instance, there is a limit 
to the swirl ratio as an overly strong swirl action has the effect of 
lowering the temperature of exhaust gas. 
Alternatively, the ignition and combustion efficiency of an air-fuel 
mixture may be improved with a tumble of a tumble ratio of greater than 
1.5. To obtain a strong tumble of a ratio above 1.5, it is desirable to 
open greatly the turbulence control valve 25 so as to allow the air-fuel 
mixture to enter the combustion chamber 2 mainly through the second intake 
port 4. In the same manner as previously described in regard to an overly 
strong swirl rate, there is also a limit to the tumble rate. 
As described above, because the first intake valve 7 is of a low-lift type, 
the speed of an air-fuel mixture entering the combustion chamber 2 through 
the first intake port 3 is increased in such a manner as to promote the 
creation of turbulence energy in the combustion chamber 2, thus raising 
the ignition and combustion efficiency of the air-fuel mixture. It is of 
course that the second intake valve 8 of a low-lift type increases the 
speed of an air-fuel mixture and promotes greatly the ignition and 
combustion efficiency in the same manner as stated above. 
In the exhaust gas purifying system of the present invention which utilizes 
an air-mixing type of fuel injector 24, when the engine is started at cold 
temperatures, mixing air is supplied to the fuel injector 24 to promote 
atomization and vaporization of an air-fuel mixture in order to increase 
the ignition and combustion efficiency of the air-fuel mixture during the 
period of time for which acceleration of a rise in the temperature of 
exhaust gas is made. This also has the effect of lowering the rate of 
hydrocarbon (HC) emissions in the exhaust gas. It is to be noted that the 
engine CE equipped with the exhaust gas purifying system of the present 
invention stop the supply of mixing air during ordinarily idling as well 
as conventional engines utilizing air-mixing type fuel injectors. Further, 
in the exhaust gas purifying system of the present invention, upon a cold 
engine start, the energy supplied to the spark plug 12 is increased by 
controlling the ignition mechanism F during the period of time for which 
acceleration of a rise in the temperature of exhaust gas is made, with the 
effect of improving the ignition and combustion efficiency of the air-fuel 
mixture. Resultingly, it becomes possible to maintain cycling stability of 
engine CE and suppress torque fluctuations even if the ignition timing is 
retarded in a range significantly after top-dead-center. 
The following discussion will be directed to a practical example in which 
the effect of swirl turbulence is utilized as a means of increasing the 
ignition and combustion efficiency of an air-fuel mixture. 
FIG. 10 shows changes in average effective pressure (Pi), i.e. torque 
changes, in relation to ignition timing under a condition where a swirl is 
present and a condition where a swirl is absent. As FIG. 10 demonstrates, 
there is a significant decrease in average effective pressure (Pi) in the 
case where a swirl is present as compared to the case where a swirl is 
absent. This indicates that, even when the ignition timing is 
significantly retarded to a point after top-dead-center, it is ensured to 
stabilize engine cycling and suppress torque fluctuations by promoting and 
inducing the creation of a swirl in the combustion process. 
In further regard to a swirl, FIG. 11 shows an example of a change in 
in-cylinder turbulence energy of an air-fuel mixture in relation to crank 
angle. As FIG. 11 demonstrates, turbulence energy is markedly induced 
after top-dead-center (360.degree. in crank angle), even at a crank angle 
of 90.degree. after top-dead-center. This turbulence energy has the effect 
of increasing the ignition and combustion efficiency of the air-fuel 
mixture. 
FIGS. 12(A) and 12(B) show the distribution patterns of an fuel-air 
mixture, with and without swirl turbulence, in the combustion chamber 2 
when the piston is at bottom-dead-center (BDC). As FIGS. 12(b) and 12(a) 
demonstrate, the mixture density distribution pattern is completely 
uniform when a swirl is present, while the absence of a swirl causes three 
different mixture densities consisting of a rich range R1, a moderate 
range R2, and a lean range R3. Accordingly, the presence of a swirl 
induces a uniform mixture distribution which in turn aids in the efficient 
ignition and combustion of the air-fuel mixture. 
Upon a cold engine start, the in-cylinder air-fuel ratio is set within a 
range from 13.5 to 18.0 during the period of time for which acceleration 
of a rise in the temperature of exhaust gas is made. In this instance, it 
is more desirable to set an excessive air ratio .lambda. at more than 1, 
in other words, to set an in-cylinder air-fuel ratio (A/F) at greater than 
14.7. Thus setting the in-cylinder air-fuel ratio near the stoichiometric 
air-fuel ratio of 14.7 or slightly on the lean side provides the effect of 
raising the temperature of exhaust gas. When setting the in-cylinder 
air-fuel ratio in a range from 13.5 to 18.0 reduces the amount of latent 
heat necessary to atomize fuel or the amount of developed or actualized 
heat necessary to raise fuel temperature, as compared to the in-cylinder 
air-fuel ratio in a range from 12 to 13 conventionally utilized during a 
cold start. The saved part of heat has the effect of raising the 
temperature of exhaust gas. Together, because of the existence of adequate 
oxygen in the combustion chamber 2, the amount of heat generated from more 
complete combustion of the air-fuel mixture is increased accompanied by an 
effective rise in exhaust gas temperature. As a result, the exhaust 
purifying catalyst is more quickly brought to its effective activation 
temperature. Moreover, the in-cylinder air-fuel ratio is set on the leaner 
side more in comparison to the conventional engines, the production of 
hydrocarbons (HC) is lowered and emission control performance is enhanced 
during a cold start. In this instance, when the amount of heat necessary 
to atomize fuel or raise fuel temperature is reduced by setting the 
excessive air ratio above 1, that is, by setting the in-cylinder air-fuel 
ratio on the lean side or larger than the stoichiometric air-fuel ratio, 
fuel is burned more completely with a resultant rise in exhaust gas 
temperature. As a result, the exhaust purifying catalyst is more quickly 
brought to its effective activation temperature, and the production of 
hydrocarbons (HC) is lowered and emission control performance is enhanced 
during a cold start. 
Referring to FIG. 13 which is a graph showing the change in exhaust gas 
temperature in relation to air-fuel ratio and ignition timing, a 
pronounced increase in exhaust gas temperature is obtained when an 
air-fuel ratio of 14.7 is utilized as opposed to the 13.0 air-fuel ratio 
used in the conventional engines. Further, when an air-fuel ratio of 14.7 
is utilized, compared to the 13.0 air-fuel ratio employed by the 
conventional engines, a rising rate of the temperature of exhaust gas 
markedly increases as the ignition timing is retarded. That is, as 
compared to the conventional engines, the rising rate of exhaust gas 
temperature is significantly increased. This indicates that the exhaust 
gas purifying system of the invention enables the exhaust gas purifying 
catalyst to quickly attain a temperature above the activation temperature 
through the mutual effect of making an air-fuel ratio leaner as compared 
to the conventional engines and the retardation of ignition timing. 
The following can be surmised as the cause of the mutual effect of the lean 
air-fuel ratio and retarded ignition timing. 
Generally, combustion of an air-fuel mixture in an expansion stroke and in 
an exhaust stroke is promulgated more as the temperature of exhaust gas 
rises. Accordingly, after-combustion of the air-fuel mixture is 
promulgated if the exhaust gas temperature is increased by means of 
retarding the ignition timing. However, because after-combustion will not 
occur if there is insufficient oxygen, a rich air-fuel mixture will 
decline in after-combustion effect which results from an increase in 
exhaust gas temperature. On the other hand, if an air-fuel mixture of an 
air-fuel ratio of greater than 14.7 is provided, a sufficient amount of 
oxygen is present, and this leads to adequate inducement of the 
after-combustion effect due to the increased exhaust gas temperature. 
Resultingly, since the after-combustion effect causes an increase in 
exhaust gas temperature, the exhaust gas temperature can be markedly 
increased when the lean air-fuel mixture and retarded ignition timing are 
concurrently employed. An air-fuel ratio range where the excessive air 
ratio .lambda. is equal to or greater than 1 have the most pronounced 
effect on raising exhaust gas temperature. 
In regard to the conventional engines, because a cold start of the engine 
shows aggravation of fuel atomization and vaporization and lowers the 
efficiency of ignition and combustion of an air-fuel mixture, the air-fuel 
mixture is enriched, for example, within a range of air-fuel ratio from 11 
to 13 which is greatly rich as compared to the stoichiometric air-fuel 
ratio. Contradistinctively, as described previously, the exhaust gas 
purifying system of the present invention employs a swirl, a tumble and 
assist air to form a desired fuel-air mixture and promote more efficient 
combustion of the air-fuel mixture, thus allowing the air-fuel ratio in 
the range of 13.5 to 18.0 to be utilized during a cold engine start, that 
is to say, efficient ignition and combustion of the air-fuel mixture can 
be maintained even with an excessive air ratio .lambda. equal to or 
greater than 1. 
It is desirable during a cold engine start and the period of time for which 
acceleration of a rise in the temperature of exhaust gas is made to supply 
secondary air to the exhaust duct 60 upstream from the catalytic converter 
45 so as to provide an exhaust air-fuel ratio of more than 14.5. In this 
instance, the secondary air supply, concurrently with the rise in exhaust 
gas temperature promulgated by the retardation of ignition timing to a 
point after top-dead-center, acts to combust the residual fuel in the 
exhaust gas and yield further a raise in the temperature of exhaust gas 
due to heat from the combustion. Resultingly, heating of the exhaust gas 
purifying catalyst is promoted and the effective activation temperature is 
attained quickly. Because an exhaust air-fuel ratio of more than 14.5 is 
employed, the residual fuel in the exhaust gas is certainly combusted, so 
as to rise the temperature of exhaust gas, thereby quickly attaining the 
effective activation temperature. 
FIG. 14 shows the temperature of exhaust gas in relation to the exhaust 
air-fuel ratio, when secondary air is supplied to the exhaust duct 60 
upstream from the catalytic converter 45. As FIG. 14 demonstrates, the 
supply of secondary air increases the temperature of exhaust gas when an 
air-fuel mixture is richer than the stoichiometric air-fuel mixture. 
However, when the exhaust gas remains a same exhaust air-fuel ratio, the 
leaner the in-cylinder air-fuel ratio is, the higher the temperature of 
exhaust gas becomes. That is, when the total amount of air is the same, 
supplying the entire air to the combustion chamber 2 results in a greater 
increase in the temperature of exhaust gas than if part of that air is 
supplied as secondary air to the exhaust gas. 
It is desirable for the catalyst activation control to maintain engine 
speed at a specified relatively high level, 2,000 rpm for example, by 
means of regulating the amount of bypass air through the idle speed 
control (ISC) valve 22 and air control valve 23 during the period of time 
for which acceleration of a rise in the temperature of exhaust gas is made 
while the engine is running in a cold condition. This causes the amount of 
heat per unit time generated in combustion chamber 2 to increase with the 
result that the temperature of exhaust gas is more increased, so that the 
exhaust purifying catalyst attains its effective activation temperature 
extremely quickly. 
As was described previously, the characteristic of change in exhaust gas 
temperature in relation to ignition timing for various engine speeds and 
the characteristic of change in estimated catalyst temperature, 20 seconds 
after an initial engine start, in relating to ignition timing are shown in 
FIGS. 8 and 9, respectively. In this instance, while the temperature of 
exhaust gas rises as the ignition timing is retarded, the time at which 
the temperature of exhaust gas begins to rise suddenly is directly related 
to the increase in engine speed. This results from the fact that, as the 
delay of ignition is defined by a time after discharge of the spark from 
spark plug 12 and the time of ignition is defined by a crank angle, a 
large change in crank angle occurs at higher engine speeds during the 
delay time of ignition even though which is unchanged. For this reason, 
when the engine operates at high speeds, the effect of raising the 
temperature o exhaust gas is adequately obtained even when the ignition 
timing is relatively advanced. 
In the exhaust gas purifying system, the engine speed may be lowered by 
once again retarding the ignition timing after a certain time interval has 
passed during the period of time for which acceleration of a rise in the 
temperature of exhaust gas is made. This reduction in engine speed reduces 
engine noise as well as a yank to a start when the vehicle is put into 
gear, and improves the product appeal of the engine CE itself, and hence, 
of the vehicle into which the engine CE is installed. 
Alternatively, after a certain time interval has passed during the period 
of time for which acceleration of a rise in the temperature of exhaust gas 
is made, the engine speed may lowered through a reduction in the amount of 
bypass air. In this case, the same effect is obtained. The retardation of 
ignition timing and reduction in the amount of bypass air may both be 
employed to lower the engine speed. 
The catalyst activation control returns the ignition timing to a normally 
advanced timing when the vehicle starts during the period of time for 
which acceleration of a rise in the temperature of exhaust gas is made. 
Returning the ignition timing increases engine output torque, resulting in 
an improvement of the starting and running performance of the vehicle in 
which the engine EC is installed, which leads to the favorable 
marketability of the vehicle. Otherwise, it is also preferable that the 
supply of bypass air may be gradually constricted and cut off after the 
vehicle starts in order to prevent a sudden decrease in the amount of 
intake air upon a start and thereby to improve the starting and running 
performance of the vehicle. It is further preferable to gradually cut off 
the supply of bypass air in order to prevent a sudden decrease in the 
amount of intake air upon a start and thereby to improve the starting and 
running performance of the vehicle. 
The catalyst activation control may set the ignition timing retarded from a 
normally advanced timing before top-dead-center when the vehicle starts 
during the period of time for which acceleration of a rise in the 
temperature of exhaust gas is made. This increases engine output torque, 
resulting in an improvement of the starting and running performance of the 
vehicle in which the engine EC is installed, which leads to the favorable 
marketability of the vehicle. The ignition timing may be left retarded in 
order to promote the rise of exhaust gas temperature, when the engine CE 
operates at a level of output torque or engine load less than a 
predetermined level, that is, in a range of operating conditions where the 
engine CE needs output torque at a low level but higher than output torque 
necessary to idle. In order for the engine CE to have cycling stability 
after the vehicle starts, the catalyst activation control may reduce a 
rate of ignition timing retardation as the engine output torque or engine 
load increases. 
In cases where the vehicle starts and then stops, if the effective 
activation temperature of the catalyst element of the catalytic converter 
45 has not yet been attained, it is preferred that, while shaping an 
air-fuel mixture and promoting combustion for maximum thermal discharge, 
the catalyst activation control once again reset the ignition timing 
retarded to a point after top-dead-center. Thus, even with the vehicle 
coming to a halt, while combustibility of the air-fuel mixture and engine 
cycling stability at the necessary minimum level are maintained, the time 
needed to attain the effective activation temperature is shortened. 
Running or stopping of the vehicle may be determined, for instance, based 
on transmission positions or vehicle speeds. Specifically, when the 
transmission is in a neutral position, it is determined that the vehicle 
is stopping. Otherwise, when a vehicle speed sensor detects a speed less 
than, for instance, 5 Km/h, it may be determined that the vehicle is 
stopping. 
During the period of time for which acceleration of a rise in the 
temperature of exhaust gas is made after a cold engine start, it is 
desirable for the catalyst activation control to set the ignition timing 
to a point after top-dead-center after the engine CE boosts its speed 
sharply up to, for instance, approximately 2,000 rpm following a full 
explosion, at which the engine speed is between approximately 500 and 600 
rpm slightly less than an engine idle speed, immediately after cranking. 
In this manner, the engine CE has been sufficiently stable in cycling 
until the time the ignition timing is set to a point after 
top-dead-center, yielding reduced torque fluctuations during the 
acceleration of a rise in the temperature of exhaust gas and improving the 
marketability of the engine CE. It is more desirable for the catalyst 
activation control to set the ignition timing to a point after 
top-dead-center after a certain time interval from a boost of engine speed 
following a full explosion of the engine after cranking. This increases 
the stability of engine cycling, leads to reducing torque fluctuations 
during the period of time for which acceleration of a rise in the 
temperature of exhaust gas is made, which always improves the 
marketability of the engine CE. Together, the catalyst activation control 
may set the ignition timing to a point advanced from the ignition timing 
in normal driving conditions, until a full explosion is completed after 
engine cranking. This advanced setting yields the effect of improving the 
ignition and combustion efficiency of an air-fuel mixture, quickly 
bringing the engine CE into self-cycling and, as a consequence, improving 
engine starting performance. 
The catalyst activation control may be simply and completely terminated by 
means of a timer after a certain time interval from the commencement of 
the period of time for which acceleration of a rise in the temperature of 
exhaust gas is made after a cold start. This simplifies the engine control 
system and/or the control logic employed and reduces manufacturing costs. 
In this regard, the period of time for which acceleration of a rise in the 
temperature of exhaust gas is made may be terminated at attainment of a 
specific temperature of the exhaust gas measured at a point behind the 
catalytic converter 45. This results in raising the temperature of exhaust 
gas until the catalyst becomes reliably active and promoting the 
activation of the catalyst. In addition, this allows the net output of 
engine to be quickly increased after the achievement of activation of the 
catalyst and yields the stability of engine cycling. 
FIGS. 17(A) through 17(D) show one example of changes in engine speed, 
intake air amount, ignition timing, and air-fuel ratio, respectively, on 
the passage of time while the catalyst activation control is conducted 
after a cold start. In this example, the ignition timing is advanced at a 
crank angle of approximately 7.degree. before top-dead-center until a 
boost of engine speed to approximately 2,000 rpm and retarded at a crank 
angle of approximately 15.degree. after top-dead-center after the boost of 
engine speed to approximately 2,000 rpm. Together, the amount of intake 
air and ignition timing are kept at steady values during the period of 
time for which acceleration of a rise in the temperature of exhaust gas is 
made so as to maintain the engine speed at approximately 2,000. 
FIGS. 18(A) through 18(C) show another example of changes in engine speed, 
ignition timing, and air-fuel ratio, respectively, on the passage of time 
while the catalyst activation control is conducted after a cold start. 
This example shows the case where the ignition timing is advanced at a 
crank angle of approximately 7.degree. before top-dead-center until a 
boost of engine speed to approximately 2,000 rpm and retarded at a crank 
angle of approximately 15.degree. after top-dead-center at the time of the 
boost of engine speed to approximately 2,000 rpm, and further retarded 
after a moment subsequent to the first retardation with the result of a 
fall of engine speed to approximately 1,500 rpm. 
FIGS. 19(A) shows still another example of changes in engine speed and 
ignition timing on the passage of time while the catalyst activation 
control is conducted after a cold start. In this example, the ignition 
timing takes an initial advanced angle i.sub.3, which is relatively large, 
until an explosion time t.sub.1 the engine CE makes a full explosion and 
attains a speed n.sub.1 (which is referred to as an explosion speed) 
slightly lower than an idle speed n.sub.2, and a normal advanced angle 
i.sub.2, which is set according to an engine speed and the amount of 
intake air and relatively small, during a time interval (d.sub.1) from the 
explosion time t.sub.1 to a time t.sub.2 the engine CE boosts its speed to 
a maximum speed n.sub.3 (which is referred to as a boost time). After the 
boost of engine speed to approximately 2,000 rpm, the ignition timing is 
further retarded to a specified crank angle i.sub.1 after top-dead-center. 
For a certain time interval d.sub.2 from the boost time t.sub.2 to a time 
t.sub.3, the retardation of the ignition timing to the specified crank 
angle i.sub.1 after top-dead-center may be delayed so as to improve the 
stability of engine cycling. Further, as shown in FIG. 19(A), if the boost 
of engine speed is detected by monitoring a change in engine speed, the 
ignition timing may be retarded to the specified crank angle i.sub.1 when 
the maximum engine speed n.sub.3 is monitored at a moment t.sub.2 '. In 
this case, it is desirable to increase the amount of intake air by, for 
instance, increasing the opening of the idle speed control (ISC) valve 22 
during the time interval d.sub.1 and, after the time interval d.sub.1, to 
further increase the amount of intake air. This enables the engine CE to 
boost its speed smoothly or rapidly. Otherwise, in cases where the 
air-fuel ratio is fixed, it is desirable to increase the opening of the 
idle speed control (ISC) valve 22 so as to increase the amount of intake 
air with an increase in retardation of the ignition timing. Increasing the 
amount of intake air results in increasing the amount of fuel to an 
air-fuel mixture and consequently, in increasing heat from combustion of 
the air-fuel mixture, leading to an accelerated rise in the temperature of 
exhaust gas. 
The ignition timing may further be changed as shown in FIG. 19(B). In this 
example, the ignition timing takes an initial advanced angle i.sub.3 until 
the explosion speed n.sub.1 is attained at an explosion time t.sub.1, and 
an angle i.sub.2 further advanced from the initial angle i.sub.3 during a 
time interval (d.sub.1) from the explosion time t.sub.1 to a boost time 
t.sub.2. After the boost of engine speed, the ignition timing is retarded 
to a specified crank angle i.sub.1 after top-dead-center. For a certain 
time interval d.sub.2 from the boost time t.sub.2 to a time t.sub.3, the 
retardation of the ignition timing to the specified crank angle i.sub.1 
after top-dead-center may be delayed so as to improve the stabilize of 
engine cycling. In this instance, fixing the amount of intake air provides 
the enhanced stability of engine cycling. 
FIG. 20 shows an example of a change in the temperature of exhaust gas on 
the passage of time (which is represented by a characteristic curve G1) 
when the catalyst activation control is performed during a cold engine 
start, in comparison with a conventional art change in the temperature of 
exhaust gas on the passage of time (which is represented by a 
characteristic curve G2). FIG. 21 shows an example of a change in the 
efficiency of purifying hydrocarbons (HC) (which is represented by a 
characteristic curve H1) as it occurs after a cold engine start when 
employing the catalyst activation control, in comparison with the prior 
art efficiency of purifying hydrocarbons (HC) (which is represented by a 
characteristic curve H2). 
As FIGS. 20 and 21 demonstrate, the engine CE employing the catalyst 
activation control realizes a significantly rapid rise in the temperature 
of exhaust gas as compared to the prior art engines. Resultingly, the 
hydrocarbon purification rate is markedly increased to a level of 
approximately 88% 20 seconds after an engine start before the vehicle 
starts. 
As described above, according to the aspects of the present invention, 
without the need for catalyst heating devices such as an EHC or an 
afterburner, while stabilizing engine cycling to suppress torque 
fluctuations during a cold engine start, the temperature of the exhaust 
gas is quickly raised, and the effective activation temperature of the 
exhaust gas purifying catalytic element is quickly attained when the 
engine is running in a non-warmed up condition. 
The system that Japanese Unexamined Patent Publication No. 2-64253 teaches 
increases the amount of intake air and retards an ignition timing during a 
cold engine start with the purpose of stabilizing an idle speed of the 
engine and rapidly warming up the engine. While, the system of the 
Japanese Unexamined Patent Publication No. 2-64253 is the same as that of 
the present invention in regard to retarding the ignition timing, 
nevertheless, the system does not incorporate a catalytic converter nor 
retards the ignition timing to a point after top-dead-center. Accordingly, 
the prior art system is of different construction and operation and does 
not achieve the effect that the present invention provides. 
FIG. 22 shows the experimental result relating to the dependency of the 
temperature of exhaust gas on ignition timing and swirl ratio and the 
dependency of time necessary to combust an air-fuel mixture on ignition 
timing and swirl ratio. As FIG. 22 demonstrates, the retardation of 
ignition timing causes a progressive rise in the temperature of exhaust 
gas. This proves the significant effect of the catalyst activation control 
employed in the present invention in which the ignition timing is retarded 
in order to cause a rise in the temperature of exhaust gas. Together, as 
FIG. 22 demonstrates, the combustion time is shortened with the effect of 
improvement of the stability of combustion of an air-fuel mixture by means 
of inducing a strong swirl of, for instance, a swirl ratio of 3.0 into the 
air-fuel mixture. Accordingly, in the catalyst activation control in which 
an enhanced swirl is generated, possible aggravation of the stability of 
combustion of an air-fuel mixture resulting from the retardation of 
ignition timing is compensated by the enhanced swirl and the substantial 
stability of combustion is maintained, consequently. 
FIG. 23 shows the experimental result relating to the dependency of a 
change in average effective pressure (Pi) on the temperature of exhaust 
gas and swirl ratio. As FIG. 23 demonstrates, the change in average 
effective pressure (Pi), on one hand, increases with a rise in the 
temperature of exhaust gas and, the other hand, significantly decreases 
with an increase in swirl ratio. This proves that, while, in the catalyst 
activation control in which an enhanced swirl is generated, the change in 
average effective pressure (Pi) tends to increase as the ignition timing 
is retarded, the enhanced swirl has an effect of suppressing or 
compensating the increasing change in average effective pressure (Pi). As 
a result, the stability of combustion and engine cycling is improved even 
when the ignition timing is retarded. 
FIG. 24 shows the experimental result relating to the dependency of a lean 
limit to air-fuel ratio on ignition timing and swirl ratio. As FIG. 22 
demonstrates, enhancing a swirl has an effect of shifting the lean limit 
of air fuel ratio to the leaner side. While, in general, the lean limit of 
air fuel ratio shifts to a less lean side as the ignition timing is 
retarded, however, in the catalyst activation control of the present 
invention in which an enhanced swirl is generated, the lean limit of air 
fuel ratio is on a more leaner side. Resultingly, the stability of 
combustion and engine cycling is improved even when the ignition timing is 
retarded. 
Experiments were performed by the applicant of this application to find the 
dependency of hydrocarbon concentration on the temperature of exhaust gas 
or crank angle. In this instance, the term "hydrocarbon (HC) concentration 
index" used herein shall mean and refer to a parentage of hydrocarbon (HC) 
concentration in exhaust gas relative to hydrocarbon (HC) concentration 
which will be obtained at the optimum ignition timing in order for the 
engine CE to provide maximum output torque. More specifically, the 
hydrocarbon concentration index is a percentage of hydrocarbon (HC) 
concentration at an ignition timing retarded from the optimum ignition 
timing relative to the hydrocarbon (HC) concentration at the optimum 
ignition timing which is defined as 100%. 
FIG. 25 shows the experimental result relating to the dependency of a 
hydrocarbon concentration index on the temperature of exhaust gas and 
swirl ration. As FIG. 25 demonstrates, the hydrocarbon concentration index 
becomes small with a rise in the temperature of exhaust gas and, is, 
however, independent of swirl ratios. Consequently, in the catalyst 
activation control of the present invention in which a rise in the 
temperature of exhaust gas is promoted by retarding the ignition time, the 
emission of hydrocarbons (HC) is significantly reduced. 
FIG. 26 shows the experimental result relating to the dependency of the 
hydrocarbons (HC) concentration index in exhaust gas on crank angle, at 
which combustion of 90% of fuel (which is referred to as 90% combustion 
crank angle) is achieved, and swirl ratio. As FIG. 26 demonstrates, making 
the 90% combustion crank angle larger, in other words, setting the 
retardation of ignition timing larger, has an effect of lowering the 
hydrocarbon concentration index and is, however, independent of swirl 
ratios. Consequently, in the catalyst activation control of the present 
invention in which a rise in the temperature of exhaust gas is promoted by 
retarding the ignition time, the emission of hydrocarbons (HC) is 
significantly reduced. 
Further, experiments were performed by the applicant of this application to 
find the dependency of carbon monoxide concentration on the temperature of 
exhaust gas. In this instance, the term "carbon monoxide (CO) 
concentration index" used herein shall mean and refer to a parentage of 
carbon monoxide (CO) concentration in exhaust gas relative to carbon 
monoxide (CO) concentration which will be obtained at the optimum ignition 
timing in order for the engine CE to provide maximum output torque. More 
specifically, the hydrocarbon concentration index is a percentage of 
hydrocarbon (HC) concentration at an ignition timing retarded from the 
optimum ignition timing relative to the carbon monoxide (CO) concentration 
at the optimum ignition timing which is defined as 100%. 
FIG. 27 shows the experimental result relating to the dependency of the 
carbon monoxide concentration index on the temperature of exhaust gas and 
swirl ratio. As FIG. 27 demonstrates, while the carbon monoxide 
concentration index is almost independent of the temperature of exhaust 
gas, it is significantly small when a swirl has a large swirl ratio. 
Consequently, in the catalyst activation control of the present invention 
in which an enhanced swirl is generated, the emission of hydrocarbons (HC) 
is significantly reduced. 
It is to be understood that although the present invention has been 
described in connection of preferred embodiments thereof, various other 
embodiments and variants may occur to those skilled in the art, which are 
within the scope and spirit of the invention, such other embodiments and 
variants are intended to be covered by the following claims.