Note: Descriptions are shown in the official language in which they were submitted.
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DESCRIPTION OF THE INVENTION
The present invention relates to a method o active
thermoatmosphere combustion and to an internal combustion
engine of an active thermoatmosphere combustion type.
With regard to an internal combustion engine, for
example, a 2-cycle engine, it has heen known tha-t self
ignition of the fresh combustible mixture can be caused in
the combustion chamber of an engine wi-thout the fresh
combustible mixture being ignited by the spark plug. The
combustion caused by the above-mentioned self ignition is
conventionally called an extraordinary combus-tion or a run
on. In the attached Figs. 1 and 2, A shows the region of
occurrence of the extraordinary combustion which is caused
in a 2-cycle engine. In Fig. 1, the ordinate indicates a
delivery ratio DR and the abscissa indicates an air-fuel
ratio A/F. On the other hand, in Fig. 2, the ordinate
indicates a delivery ratio DR and the abscissa indicates
the number of revolutions per minute N of the engine. In
addition, Fig. 1 shows the results of an experiment
conducted under a constant engine speed of 2000 r.p.m. and
Fig. 2 shows the results of an experiment conducted under
a constant air-fuel ratio of 15:1.
In a 2-cycle engine, when the engine is operating
at a high speed under a light load, wherein the above-
-mentioned extraordinary combus-tion is caused, the amount
of residual exhaust gas remaining in the cylinder of the
engine is much larger than that of the fresh combustible
mixture fed into the cylinder. Therefore, the fresh
combustible mixture fed into the cylinder is heated until
it is reformed by the residual exhaus-t gas, which has a
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high temperature, and as a result, the fresh combustible
mixture produces radicals. An atmosphere wherein radicals
are produced as mentioned above is hereinaEter called an
active thermoatmosphere. However, when an extraordinary
combustion is caused, the ac-tive thermoatmosphere is
extinguished at the beginning of the compression stroke,
and a hot spot ignition, a mis-fire and an explosive
combustion caused by a spark plug are alternately repeated,
thus, causing a great fluctuation of torque. Since -the
extraordinary combustion has drawbacks in tha-t a great
fluctuation of torque occurs as mentioned above and, in
addition, the piston will melt due to the occurrence of
the above-mentioned hot spot ignition, such an extraordinary
combustion is conventionally considered an undesirable
combustion.
The inventor conducted research on ex-traordinary
combustion and, as a result, has proven that, if the
active thermoatmosphere which is caused in the extraordi-
nary combustion at the beginning of the compression stroke
can continue to be maintained until the end of the compression
stroke, self ignition of the active thermoatmosphere is
caused in the combustion chamber of an engine wi-thout the
thermoatmosphere being ignited by the spark plug and,
then, the active thermoatmosphere combustion takes place.
In addition, the inventor has further proven that -this
active thermoa-tmosphere combustion results in quiet engine
operation and can be caused even iE a lean air-fuel
mix-ture is used. This results in a considerable improvement
in fuel consumption and a considerable reduc-tion in the
amoun-t of harmful components in the exhaust gas. Particularly
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- in an engine for use in, for example, a vehicle, the
majority of the opera-tion of the engine is carried out
under a partial load. Consequently, if -the above-mentioned
active thermoatmosphere combustion is carried out under a
partial load, the fuel consumption is considerably improved
and the amount of harmful components is considerably
reduced.
An object of the present invention is to provide an
internal combustion engine and a method of operation
thereof which are capable of always creating a stable
active thermoatmosphere independent of the number of
revolutions per minute of the engine when the engine is
operating under a partial load.
According to the present invention, there is
provided a method of c~mbustion in an internal combustion
engine having a combustion chamber therein, said method
comprising the steps of:
restricting a fresh combustible mixture fed into
the combustion chamber when the engine is operating under
a partial load for maintaining the flow velocity of the
combustible mixture flowing into the combustion chamber at
a low level;
feeding the combustible mixture into the combustion
chamber while suppressing the flow and turbulence of the
burned ga-s in the combustion chambèr and preventing the
dissipation of the hea-t of the burned gas contained in the
combustion chamber for'maintaining the residual burned gas
in the combustion chamber at a high temperature;
creating an active thermoatmosphere in the
combustion chamber at the beginning of the compression
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stroke;
con-tinuing to maintain the active thermoatmos-
phere until the end of the compression stroke and reforming
the fresh combustible mixture, and;
causing a self-ignition of the fresh combustible
mixture.
In addition, according to the present invention, there
is provided a 2-cycle internal combustion engine comprising:
a cylinder block having a cylinder bore and a
crank room therein;
a piston having an approximately flat top
surface and reciprocally moving in said cylinder bore,
said piston and said cylinder bore defining a combustion
chamber;
a scavenging passage having a scavenging port
which opens into said combustion chamber and communicating
said combustion chamber with said crank room for feeding a
fresh combustible mixture into said combustion chamber,
said scavenging port being arranged to be directed to an
approximately central portion of combustion chamber;
an exhaust passage having an exhaust port
opening into said combustion chamber for discharging the
burned gas from said combustion chamber into the atmosphere,
and;
restricting means disposed in said scavenging
passage at a position near the position wherein sald
scavenging passage opens into said crank room for
restricting the flow velocity of the fresh combustible
mixture fed into said combustion chamber when the engine
is operating under a partial load and for creating an
- active thermoatmosphere in said combustion chamber to
cause the self ignition of the fresh combustible mixture.
.. .. . ..
Furthermore, according to -the present invention,
there is provided an internal combustion engine of -the
type wherein the intake stroke is started after the exhaust
stroke'is completed, said engine comprising:
a housing having a bore therein;
a piston movable in said bore, said piston and
said bore defining at least one combustion chamber;
an intake passage communicating said combustion
chamber with the atmosphere for feeding a fresh combustibIe
mixture into said combustion chamber;
an exhaust passage communicating said combustion
chamber with -the atmosphere for discharging the burned gas
into the atmosphere, and;
restricting means in said intake passage for
reducing the flow velocity of the fresh combustible mixture
fed into the combustion chamber, which is smaller than the
flow velocity in an ordinary engine, when the engine is
operating under a partial load.
The present invention may be more fully understood
from the following description of preferred embodiments of
the invention, together with the accompanying drawings.
BRIEF DESCRIPTION OF THE DRAWINGS
In the drawings:
Figs, 1 and 2 are graphs showing the region of
the occurrence of the active thermoatmosphere combustion;
Fig. 3 is a cross-sectional side view of an
embodiment of a 2-cycle e,ngine according to the present
invention;
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Fig. 4 is a cross-sectional view -taken along the
line IV-IV in Fig. 3;
Fig. 5 is a cross-sectional view taken along the
line V-V in Fig. 4;
Fig. 6 is a graph showing the change in the
opening area of the scavenging control valve and the
exhaus-t control valve in the engine shown in Fig. 3;
Fig. 7a is a diagram showing the scavenging and
exhaust strokes of the engine shown in Fig. 3;
Fig. 7b is a graph showing the veloci-ty of the
fresh combustible mixture flowing into the combustion
chamber from the scavenging port and showing the actual
scavenging timing caused by the fresh combustible mixture
in the engine shown in Fig. 3;
Fig. 8 is a cross-sectional side view of another
embodiment of a 2-cycle engine according to the present
invention
Fig. 9 is a graph showing the specific fuel
consumption and the concentrations of HC and NOX in the
exhaust gas in the engine shown in Fig. 8;
Fig. 10 is a graph showing the specific fuel
consumption in the engine shown in Fig. 8;
Fig. 11 is a cross-sectional side view of a
further embodiment of a 2-cycle engine according to the
present invention;
Fig. 12 is a cross-sec-tional side view of an
embodiment oE a 4-cycle engine according to the present
invention;
Fig. 13 is a cross-sectional side view of an
another embodimen-t of a 4-cycle enqine according to the
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present invention;
Fig. 14 is a cross-sectional side view of a
further embodiment of a 4-cycle engine according to the
present invention;
Fig. 15 is a graph showing changes in the ~pening
areas of the flow control valve, the exhaust control valve
and the throttle valve;
Fig. 16 is a cross-sectional side view of an
embodiment of a rotary-piston type engine according to the
present invention, and;
Fig. 17 is a cross-sectional side view of an
alternative embodiment of a rotary piston type engine
according to the present invention.
DESCRIPTION OF P~EFERRED EMBODIMENTS
Figs. 3 and 4 show the case wherein the present
invention is applied to a Schn~rle type 2-cycle engine.
In Figs. 3 and 4, 1 designates a cylinder block, 2 a
cylinder head fixed onto the cylinder block 1, 4 a piston
having an approximately flat top face and reciprocally
moving in a cylinder bore 3 formed in the cylinder block 1
and 5 a combustion chamber formed between the cylinder
head 2 and the piston 4; 6 designates a spark plug, 7 a
crank case, 8 a crank room formed in -the crank case 7 and
9 a balance weight; 10 designates a connecting rod, 11 an
intake pipe, 12 an intake passage and 13 a carburetor; 14
designates a throttle valve of the carburetor 13, 15 a
pair of scavenging ports, 16 a scavenging passage and 17
an exhaust port; 18 designates an exhaus-t pipe, 19 an
exhaust passage and 20 a reed valve which permits the
inflow of a Eresh combustible mixture into the crank room
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8 from the intake passage 12. The scavenging passage 16
opens into the crank room 8 at an opening 21, on one hand,
and is divided into two branches 16a, 16b which open into
the combus-tion chamber 5 at scavenging ports 15, on the
other hand. An arm 22 is fixed onto the thro-ttle valve
14, and the tip of this arm 22 is connected via a wire 23
-to an accelerator pedal 24 which is disposed in the driver's
compartment. On the other hand, a scavenging control
valve 25 is disposed in the scavenging passage 16 at a
position near the opening 21 and is fixed onto a valve
shaft 26 pivotably mounted on the cylinder block 1. A cam
27 is mounted on the valve shaft 26, and a wire 28 which
is wound on the outer periphery of the cam 27 is connected
to the accelerator pedal 24. Consequently, when the
accelerator pedal 24 is depressed, the throttle valve 14
and the scavenging control valve 25 are opened.
Fig. 6 indicates changes in opening areas of the
throttle valve 14 and the scavenging control- valve 25. In
Fig. 6, the ordinate X indicates a ratio of an opening
area to the full opening area of the scavenging con-trol
valve 25, and the abscissa Y indicates a ratio of an
opening area to the full opening area of -the throttle
valve 14. The relationship between the above-mentioned
opening area ratios of the throttle valve 14 and the
scavenging control valve 25 is shown by the curved line C
in Fig. 6. As is apparent from Fig. 6, the scavenging
control valve 25 is gradually opened and then fully opened
before the throttle valve 14 reaches a position corresponding
to the opening area ra-tio X of approximately 40 percent.
In addition, the scavenging control valve 25 remains fully
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opened when the throttle valve 14 is further opened.
Consequently, in Fig. 3, the cam 27 is connected to the
valve shaft 26 in such a manner -that the cam 27 rotates
together with the valve shaft 26 until -the time the accel-
eration pedal 24 is depressed to a particular extent and,then, when the acceleration pedal 24 is further depressed
after the scavenging control valve 25 is fully opened,
only the cam 27 rotates. As mentioned above, the scavenging
passage 16 is throttled by means of the scavenging control
valve 25 when the engine is operating under a partial
load, so that the throttling operation of the scavenging
passage 16 is strengthened as the load of the engine is
reduced.
In operation, a fresh combustible mixture introduced
into the crank room 8 from the intake passage 12 via the
reed valve 20 is compressed as the piston 4 moves downwards.
Then the fresh combustible mixture under pressure in the
crank room 8 flows into the combustion chamber 5 from the
scavenging port 15 via the scavenging passage 16 when the
piston 4 opens the scavenging port 15. At this time, if
the scavenging control valve 25 remains slightly opened,
the st~eam of the fresh combustible mixture flowing into
the combustion chamber 5 from the crank room 8 via the
scavenging passage 16 is restricted by the scavenging
control valve 25. As a result of this, the flow velocity
of the fresh combustible mixture is reduced. Since the
flow velocity of the fresh combustible mixture is low
throughout the inflow operation of the fresh combustible
mix-ture, due to the restricting operation of the scavenging
30. control valve 25, -the flow of the residual burned gas in
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the combustion chamber 5 is extremely small and, as a
result, the dissipation of the hea-t of the residual burned
gas is prevented. In addition, at the beginning of the
compression stroke under a partial load of the en~ine, a
large amount of the residual burned gas is present in the
combustion chamber 5. Since -the amount of the residual
burned gas in the combustion chamber 5 is large and, in
addition, the residual burned gas has a high temperature,
the fresh combustible mixture is heated until it is reformed
by the residual burned gas and, as a result, an active
thermoatmosphere is created in the combustion chamber 5.
Further, since the flow of the gas in the combustion
chamber 5 is extremely small during the compression stroke,
the occurrence of turbulence and the loss of heat energy
are-restricted to the smallest persible extent. Consequently
the active thermoatmosphere thus created continues to be
maintained during the compression stroke and, as a result,
-the self ignition of the active thermoatmosphere is caused
consequently, the combustion is advanced while being
controlled by the residual burned gas. As mentioned
previously, this ignition is not caused by the spark plug
6. When the piston 4 moves downwards and opens the exhaust
port 17, the burned gas in the combustion chamber 5 is
discharged into the exhaust passage 19.
As mentioned above, in order to cause the active
thermoatmosphere combustion, it is necessary to continue
to maintain the active thermoatmosphere until the end of
the compression stroke. ~lowever, it is impossible to
continue to maintain the active thermoatmosphere un-til the
end of the compression stroke by merely throttling the
scavenging passage 16 by means of the scavenging control
valve 25 disposed in the scavenging passage 16. That is,
if a flow or turbulence of the residual burned gas in the
combustion chamber 5 is caused, the heat of the residual
burned gas escapes into the cylinder wall. As a result of
this, since -the residual burned gas is cooled, i-t is
impossible to continue to maintain the active thermoatmos-
phere until the end of the compression stroke. The fresh
combustible mixture flowing into the combustion chamber 5
from the scavenging port 15- has a great influence on the
creation of the above-mentioned flow and turbulence of the
residual burned gas. According to the experiments conducted
by the inventor, it has been proven that the velocity of
the fresh combustible mixture flowing into the combustion
chamber, the inflow direction of the fresh combustible
mixture and the strength of the turbulence of the fresh
combustible mixture immediately before the fresh combustible
mixture flows into the combustion chamber, have a great
influence on the creation of the flow and the turbulence
of the residual burned gas.
Fig. 7(a) is a diagram illustrating the opening and
closing timings of the scavenging and the exhaust ports of
the 2~cycle engine shown in Fig. 3. Fig. 7(b) is a graph
wherein the ordinate indicates the velocity V of the fresh
combustible mixture flowing into the combustion chamber
from the scavenging port and the abscissa indicates the
crank angle. In Figs. 7(a) and (b), EO indicates an
opening timing of the exhaust port, SO an opening timing
of the scavenging port, SC a closing timing of the scavenging
port and EC a c:Losing timing of the exhaust port; P indicates
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a timing of the start of the inflow of the fresh comhustible
mixture into the combustion chamber from the scavenging
port and Q a timing of the completion of said inflow of
the fresh combustible mixture. In the crank angle between
SO and P in Figs. 7(a~ and (b), even if the scavenging
port is opened, since the pressure in the combustion
chamber is higher than tha-t in the scavenging passage, the
burned gas in the combustion chamber does not flows into
the scavenging passage. On the other hand, when the
piston starts the upward movement after it reaches the
bottom dead center, the pressure in the combustion chamber
is again increased. As a result, in the crank angle
between Q and SC in Figs. 7(a) and (b), since the pressure
in the combustion chamber becomes higher than that in the
scavenging passage and the gas in the combustion chamber
flows backwards into the scavenging passage. Consequently,
the fresh combustible mixture flows into the combustion
chamber from the scavenging port during the time period
shown by the hatching in Fig. 7(a). In Fig. 7(b), the
curved line E shows change in velocity V of the fresh
combustible mixture flowing into the combustion chamber
from the scavenging port in a conventional 2-cycle engine.
As is apparent from Fig. 7(b), in a conventional engine,
since the fresh combustible mixture flows into the combustion
chamber at a high speed at the beginning of the inflow
thereof, a strong turbulence and flow of the residual
burned gas in the combustion chamber are caused by -the
fresh combustible mixture and, as a resul-t, it is impossible
to continue to maintain the active thermoa-tmosphere until
the end of the compression stroke. Contrary -to this, by
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throttling the scavenging passage 16 by means oE -the
scavenging control valve 25 as indicated in Fig. 3, the
velocity V of the fresh combustible mixture flowing into
the combustion chamber 5 from the scavenging port 15
becomes low throughout the scavenging operation caused by
the fresh combustible mixture as shown by the curved line
F in Fig. 7(b). In addition, as is shown by -the point P'
in Fig. 7(b), the start of the inf:Low operation of the
fresh combustible mixture is delayed as compared to that,
shown by P, in an conventional engine. Consequently,
since the fresh combustible mixture gently flows into the
residual burned gas, it is possible to minimize the turbulence
and flow of the residual burned gas. In addition, in
order to reduce the flow velocity of fresh combustible
mixture entering into the combustion chamber 5 from the
scavenging port 15, it is preferable that the scavenging
passage 16 be so formed that the cross-sectional area of
the scavenging passage 16 is gradually increased towards
the scavenging port 15. Furthermore, the arrangement of
the scavenging control valve 25 causes a turbulence of the
fresh combustible mixture flowing in the scavenging passage
16. However, by positioning the scavenging control valve
25 at a position as remote as possible from the scavenging
port 15, that is, at a position near the opening 21, the
-turbulence created by the scavenging valve 25 is extinguished
in the scavenging passa~e 16, and in addition, an approxi-
mately laminer f:low of the fresh combustible mixture is
caused in the scavenging passage 16. As a result of this,
the fresh combustible mixture genera-ting extremely weak
turbùelence therein flows in-to the combustion chamber 5
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from the scavenging port 15.
In addition, it is necessary to construct the
scavenging port 15 so that, as is shown by the arrows G in
Fig. 4, the fresh combustible mixture flows into the
combustion chamber 5 towards an approximately central
portion of the combustion chamber S and, at the same tlme,
as is shown by the arrow G in Fig. 5, the fresh combustible
mixture flows into the combustion chamber 5 slightly
upwards. That is, if the scavenging port 15 is so constructed
that the fresh combustible mixture flows into the combustion
chamber 5 along the circumferntial wall of the cylinder as
shown by the arrow H in Fig. 4, the fresh combustible
mixture causes turbulence and flow of -the residual burned
gas prevailing on the circumferential wall of the cylinder.
As a result of this, since the heat of the residual burned
gas in the combustion chamber easily escapes into the
cylinder wall, it is difficult to continue to maintaing
the active thermoatmosphere until the end of the compression
stroke~
sy the above-mentioned preferred arrangements and
constructions of the scavenging control valve 25, the
scavenging passage 16 and the scavenging port 15, since
the fresh combustible mixture flowing in-to the combus-tion
chamber 15 from the scavenging por-t 15 does not cause
turbulence and flow of the residual burned gas and does
not disperse in the combustion chamber 5, the dissipation
of the heat of the residual burned gas is preven-ted. As a
result of this, an active thermoatmosphere continues to be
maintained until the end of the compression stroke and, in
the region of the partial load shown by B in Figs. 1 and
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2, active thermoa-tmosphere combustion is carried out.
In addition, as is shown in Fig. 3, it is preferable
that the cylinder head 2 be so constructed that an annular
squish area Z is formed between the cylinder head 2 and
the peripheral portion of -the top face of the piston 4
when the piston 4 reaches the top dead center. In this
case, the propagation of the flame created by -the self
ignition of the active thermoatmosphere is con-trolled by
the squich flow which is caused when the piston 4 reaches
the top dead center, thus preventing the occurrence of a
detonation. As a result of this, a stable active thermoat-
mosphere combustion can be carried out.
In the 2-cycle engine shown in Fig. 3, as is shown
~y the curved line C in Fig. 6, the scavenging control
valve 25 remains fully opened when the opening area ratio
Y of the throttle valve is larger than 40 percent. Conse-
quently, when the opening area ratio Y of the throttle
valve becomes larger than 40 percent, ordinary combustion
which is caused by the spark plug 6 is carried out.
As mentioned previously, in order to continue to
maintain the active thermoatmosphere until the end of the
compression stroke, it is necessary to minimize the turbalence
- and the flow of the residual burned gas in the combustion
chamber. Two causes of turnbulence and flow of the residual
burned gas are an abrup-t blowing o~f operation of the
exhaust gas discharging from -the exhaust port 17 (Fig.3)
and interference by the pulsa-ting pressure of the exhaust
gas. In order to preven-t the above-mentioned abrupt
blowing off operation and interEerence, as is shown in
Fig. 8, it is preferable that an exhaust control valve 29
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be disposed in the exhaust passage 19. The exhaust control
valve 29 is fixed onto a valve shaft 30 pivotably moun-ted
on the exhaust pipe 18, and a cam 31 is mounted on -the
valve shaft 30. Similar to the scavenging control valve 25,
a wire 32 is wound on the ou-ter periphery of the cam 31
and is connected to the accelerator pedal 24~ The relation-
ship between the opening area ratios of the exhaust control
valve 29 and the throttle valve 14 is shown by the curved
line D in Fig. 6. In addition, in order to appropriately
prevent the exhaust gas from being abruptly discharged
from the exhaust port 17, it is preferable tha-t the volume
of the exhaust passage 19 located between the exhaust
port 17 and the exhaust control valve 29 be smaller than
that of the combustion chamber 5 when the piston is positioned
at the bot-tom dead center.
Figs. 9 and 10 indicate the results of experiments
conducted by using an engine as illustrated in Fig. 8.
The engine used had a single cylinder of 372 cc and an
effective compression ratio of 7.9:1. In addition, the
experiments related to Fig. 9 were conducted under a
constant engine revolution speed of 1500 r.p.m and a
constant airfuel ra-tio of 16:1 by changing the delivery
ratio within the range of 5 through 2 percent. In Fig. 9,
the ordinate indicates specific fuel consumption be (gr/Ps-h),
concentration of IIC (ppm) and concentration of NOX (ppm),
and the abscissa :indicates a ratio OR(%) of an opening
area to the full opening area of the exhaus-t control
valve, and a delivery ratio DR(%). In addition, in Fig.
9, the curved bro]cen lines I and j indicate the specific
fuel consumption be (gr/Ps-h) and concen-tration of HC,
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respectively, in a conventional 2-cycle engine; while the
curved solid lines IC, L and M indicate the specific fuel
consumption be, concentration of NOX and concentration of
HC, respectively, in a 2-cycle engine according to the
present invention. As is apparent from Fig. 9, the specific
fuel consumption be and the concent:ration of HC in an
engine according to the present invention are considerably
reduced as the load of the engine is reduced, that is, the
delivery ratio DR is decreased as compared with those in a
conventional engine. Fig. 10 shows the specific fuel
consumption of a 2-cycle engine according to the present
invention. In Fig. 10, the ordinate indicates the mean
effective pressure Pme, and the abscissa indicates the
number of revolutions per minute of the engine N(r.p.m.).
In addition, in Fig. 10, the numerals appearing in the
graph indicate the specific fuel consumption (gr/Ps-h).
Furthermore, in Fig. 10, the region located beneath the
solid line S is the region wherein the active thermoatmosphere
combustion is carried out. As is shown in Figs. 1, 2 and
10, active thermoatmosphere combustion is carried out
under a partial load of the engine over the entire range
of the number of engine revolutions per minute and over a
wide range of the air-fuel ratio. Particulary, as is
indicated in Fig. 1, active thermoatmosphere combustion
can be carried out by using a lean air-fuel mixture having
an air-fuel ra-tio of 16 through 21:1. Consequently, there
is an advantage in that the amount of harmful E-IC, CO and
NOX components in the exhaust gas can be simultaneously
reduced.
Fig. 11 illustrates a further embodiment of a
- 18 -
2-cycle engine according to the present invention. Referring
to Fig. 11, a switching valve 34 having a through-hole
33 therein is disposed in the scavenging passage 16 at a
position near the crank room 8. A first bypass passage 35
and a second bypass passage 36 which have a cross-sectional
area smaller than that of the scavenging passage 16 are
provided in addition to the scavenging passage 16. The
switching valve 34 is fixed onto the valve shaft 37 pivotably
mounted on the cylinder block 1, and the tip of an arm 38
fixed onto the valve shaft 37 is connected to the accelerator
pedal 24 by mPans of a wire 39. When the acceleration
pedal 24 is not depressed, that is, at thè time of iding,
the through-hole 33 of the switching valve 34 is aligned
with the second bypass passage 36 as shown in Fig. 11. On
the other hand, when the acceleration pedal 24 is depressed,
the switching valve 34 rotates in the counter-clockwise
direction and, as a result, the through-hole 33 of the
switching valve 34 is aligned with the first bypass passage
35. After this, when the acceleration pedal 24 is further
depressed, the through-hole 33 of the switching valve 34
is aligned with the scavenging passage 16. When the
switchlng valve 34 is positioned in the position shown in
Fig. 11, the fresh combustible mixture in the crank room
is introduced into the scavenging passage 16 via the
through-hole 33 and the second bypass passage 36. As is
apparent from Fig. 11, the length of the second bypass
passage 36 is longer than that of the firs-t bypass passage 35.
In addition, in t:his embodiment, a 2-cycle engine has a
plurality cylinders, and the exhaust passages 19 of all of
the cylinders are interconnected to each o~ther via a
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passage 19' located upstream of the exhaust control valve 29.
As mentioned above, when the engine is operating
under a light load, as illustrated in Fig. 11, the fresh
combustible mixture in the crank room 8 is introduced into
the scavenging passage 16 via the through-hole 33 of the
switching valve 34 and the second bypass passage 36, and
then, into the combustion chamber 5 via the scavenging
port 15. Since the second bypass passage 36 has a small
cross-sectional area and a long length, the fresh combustible
mixture flowing in the second bypass passage 36 is subjected
to the resisting operation due to the second bypass passage
36 and, as a result, the fresh combustible mixture flows
into the combustion chamber 5 from the scavenging port 15
a-t a low speed similar to the case wherein, in Fig. 3, the
scavenging control valve 25 is provided. In addition,
when the engine is operating under a light load, since the
opening degree of the throttle valve 14 is small, the
vacuum level in the intake passage 12 is large. Consequently,
when the piston 4 moves upwards, a large vacuum is produced
in the crank room 8. Contrary to this, when the piston 4
moves downwards, the pressure in the crank room 8 is
elevated. Thus, the vacuum and the pressure are alternately
produced in the crank room 8. As a result of this, the
fresh combustible mixture in the second bypass passage 36
is gradually introduced into the scavenging passage 16
while reciprocally moving within the second bypass passage
36 due to the above-mentioned alternate production of the
vacuum and the pressure. Consequently, since the fresh
combustible mixture remains in the second bypass passage-
36 for a long time, while reciprocally moving in the
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second bypass passage 36, the vaporization oE the fuel in
the second bypass passage 36 is promoted. In addition,
since a part of the fresh combustible mixture introduced
into the combustion chamber 5 due -to the reciprocal movement
of the fresh combustible mixture in the second bypass
passage 36 is again sucked into the-second bypass passage
36 when the vacuum is produced in -the crank room 8, at
this time the vaporization of the fuel contained in the
fresh combustible mixture is further promoted and, at the
same time, the fresh combustible mixture is reformed due
to the heat exchanging operation between the fresh combustible
mixture and the residual burned gas in the combustion
cha-mber. By providing the bypass passage as mentioned
above, since the vaporization o the liquid fuel con-tained
in the fresh combustible mixture is promoted and, at the
same time, the fresh combustible mixture is reformed
before the fresh combustible mixture is in-troduced into
the combustion chamber 5, it is possible to easily create
an active -thermoatmosphere in the combustion chamber 5.
As a result of this, it is possible to ensure an active
thermoatmosphere combustion when the engine is operating
under a partial load.
As mentioned previously, when the accelerator pedal
24 is depressed and, thus, the level oE the load of the
engine is increased, the fresh combustible mixture is
introduced into the scavenging passage 16 via the first
bypass passage 35, which has a leng-th shorter than that of
-the second bypass passage 36. On the o-ther hand, when the
acceleration pedal 24 is further depressed and, thus, the
engine is operating under a heavy load, the fresh combustible
- 21 -
mixture in the crank room 8 is directly introduced into
the scavenging passage 16 via the through-hole 33 of the
switching valve 34. Consequently, at this time, the flow
resistance which the fresh combustible mixture is subjected
to is reduced, whereby a desired high output power of the
engine can be obtained.
Fig. 12 illustrates the case wherein the present
invention is applied to a 4-cycle engine. In Fig. 12, 40
designates a cylinder block, 41 a piston reciprocally
movable in a cylinder bore 42 formed in the cylinder block
40, 43 a cylinder head fixed into the cylinder block 40
and 44 a combustion chamber formed between the piston 41
and the cylinder head 43; 45 designates an intake port, 46
an intake valve, 47 an intake manifold and 48 a carburetor;
49 designates a throttle valve of the carburetor 48, 50 an
exhaust port, 51 an exhaust valve, 52 an exhaust manifold,
and 62 a spark plug. An arm 53 is fixed onto the throttle
valve 49, and the tip of the arm 53 is connected to the
accelerator pedal 55 via a wire 54. On the other hand, a
flow control valve 57 is disposed in an intake passage 56,
at a position located downstream of and near the throttle
valve 49, and is fixed onto a valve shaft 58 pivo-tably
mounted on the intake manifold 47. A cam 59 is mounted on
the valve shaft 58 and a wire 60, which is wound on the
outer periphery of the cam 59, is connected -to the accelerator
pedal 55. The relationship between the opening area
ratios of the throttle valve 49 and the flow control valve
57 is equal -to t:he relationship betwèen the opening area
ratio of the throttle valve and the opening area ratio of
the exhaust control valve, which relationship is shown by
22 -
~ 5 ~5
the curved line C in Fig. 6 and was hereinbefore described
with reference to Fig. 3. Consequently, the flow control
valve 57 is gradually opened and then fully opened before
the throttle valve 49 reaches a point corresponding to the
opening area ratio of approximately 40 percent. On the
other hand, the flow control valve 57 remains fully opned
when the openinc~ area ratio of the throttle valve 49 is
larger than approximately 40 percent. With the arrangement
in Fig. 12, it is preferable that the duration of valve
overlapping, wherein both the intake valve 46 and the
exhaust valve 51 remain opened at the end of the exhaust
- stroke, be longer than that in a conventional 4-cycle
engine.
When the flow control valve 57 remains slightly
opened and, thus, the engine is operating under a light
load, the vacuum level in the intake manifold 47 is consid-
erably large. On the other hand, since the pressure in
the combustion chamber 44 and in the exhaust port 50 is
larger than the atmospheric pressure at the end of -the
exhaust stroke, when the intake valve 46 is opened at the
end of the exhaust stroke, the burned gas in the combustion
chamber 44 blows back into the intake por-t 45 via the
intake valve 46. The more the timing of the opening
operation of the intake valve 46 is advanced, that is, the
more the duration of the valve overlapping is elongated,
-the more the amount of the burned gas blowing back into
-the intake port 50 is increased. As mentioned above,
since a large amount of the burned gas blows back into the
intake manifold 47 in an engine according to the present
invention, the vaporization of the liquid fuel in the
- 23 -
~L~LC!~6~
intake manifold 47 is promoted and, at the same time, the
fresh combus-tible mixture in the intake manifold 47 is
reformed. In addition, the promotion of the vaporization -
causes a uniform distribution of the fuel into the plurality
of cylinders and also causes an improvement in the respon-
siveness of the engine with respect to the depressing
operation of the accelerator pedal. If a uni~orm distribution
o~ the fuel into the plurality of cylinders can not be
obtained as in a conventional engine, the air-fuel ratio
of the fresh combustible mixture becomes irregular among
the respective cylinders. Consequently, when a lean
air-fuel mixture is used, the air-fuel ratio in one of the
plurality of cylinders becomes large and is increased
beyond the range wherein ignition can be caused. Consequently,
in a conventional engine, in order to prevent the air-fuel
ratio in all of the plurality of cylinders from being
increased beyond the range wherein the ignition can be
caused, it is necessary to use a fresh combustible mixture
having a relatively small air-fuel ratio and, as a result,
it is difficult to use a lean air-fuel mixture. Contrary
to this, in the present invention, since a uniform distri-
bution of the fuel into the plurality of cylinders can be
obtained; a lean air-fuel mixture can be used. Particularly
at the time of idling and at the time when the engine is
operating under a light load, wherein a satisfactory
vaporizating operation of fuel cannot be obtained in a
conventional engine, the distribution of the fuel into the
plurality of cylinders becomes uniform in an engine according
to the present invention. As a result of this, since a
good combustion can be obtained in all of the cylinders,
- 24 -
the fuel consumption is grea-tly improved.
As mentioned above, in order to effectively promote
the vaporization of fuel and reform the fresh combus-tible
mixture, it is preferable that a large amount of the
burned gas be caused to blow back into the in-take manifold
47 as much as possible. To ~his end, it is preferable
that an exhaust control valve 61 be disposed in an exhaust
pipe 52' as illustrated in Fig. 12. The exhaust control
valve 61 is caused to rotate by means of a wire 63 connected
to the accelerator pedal 55 in the same manner as the
exhaust control valve 29 illustrated in Fig. 8. The
relationship between the opening area ra-tios of the exhaust
control valve 61 and the throttle valve 49 is equal to the
relationship between the opening area ratio X of the
exhaust control valve and the opening area ratio Y of the
throttle valve, shown by the curved line D in Fig. 6 and
described with reference to Fig. 3. Consequently, since
the flow rate of the exhaust gas is reduced by the exhaust
control valve 61 when the exhaust control valve 61 is
slightly opened, that is when the engine is operating
under a light load, the pressure in the combustion chamber
44 at the time of valve overlapping is greater than in the
case wherein the exhaust control valve 61 is fully opened.
As a result of this, a large amount of the exhaust gas can
be caused to blow back into the intake manifold 47.
Fig. 13 illustrates another embodiment of a 4-cycle
engine in which the vaporiza-tion of fuel can be further
promoted and the fresh combustible mixture can be further
reformed. Referring -to Fig. 13, a reed valve 64 only
permitting the downward flow of -the fresh combus-tible
- 25 -
;76~;i
mixture is provided, and an accumulator 66 having a diaphragm
65 therein is provided. An inside chamber 67 of the
accumulator 66 is connec-ted via a conduit 68 to the intake
passage 56 upstream of the flow control valve 57 on one
hand, while the inside chamber 67 îs connected via a
conduit 69 to the intake passage 56 downstream of the flow
control valve 57 on the other hand. A small throttle
valve 70 is disposed in the conduit 68 and is fixed onto a
valve shaEt 71 pivotably mounted on the conduit 68. A cam
72 is mounted on the valve shaft 71 and a wire 73 connected
to the accelerator pedal 55 is wound on the outer periphery
of the cam 72. In this embodiment, the exhaust control
valve 61 is arranged in the exhaust manifold 52.
Fig. 15 shows the relationship between the opneing
area ratios of the throttle valve 49, the flow con-trol
valve 57, the exhaust control valve 61 and the small
- throttle valve 70. In Fig. 15, the abscissa X indicates
- the ratio (%) of an opening area to the full opening area
of the throttle valve 49, and the ordinate Y indicates the
ratio (~) of an opening area to the full opening area of
the flow control valve 57, the exhaust control valve 61
and the small throttle valve 70. In Fig. 15, the curved
line P indicates the relationship between the opening area
ratios of the throttle valve 49 and the exhaust control
valve 61, and the curved line Q indicates the relationship
between the opening area ratios of the throttle valve 49
- and the.small th:rottle valve 70. In addition, in Fig. 15,
the curved line R indicates the relationship between the
opening area rat.ios of the throttle valve 49 and the flow
control valve 57. The relationship between the opening
- 26 -
.
:'',
area ratios of the throttle valve 49 and the exhaust
control valve 61 which is shown by curved line P in Fig. 15,
is equal to the relationship between the opening area
ratio Y of the -throttle valve and the opening area ratio X
of the exhaust control valve, which is shown by -the curved
line D in Fig. 6. On the other hand, as is shown by the
curved line R in Fig. 15, the flow control valve 57 remains
fully closed when the opening area ratio of the -throttle
valve is less than 50 percent, while the flow control
valve 57 is rapidly opened and then fully opèned when the
opening area ratio of the throttle valve becomes larger
than 50 percent. In addition, as is shown by the curved
line Q in Fig. 15, the small throttle valve 70 is gradually
opened as the throttle valve 49 is opned, and the small
throttle valve 70 is fully opened when the flow control
valve 37 is fully opened.
As mentioned above, since the flow control valve 57
remains fully closed when the engine is operating under a
partial load, the fresh combustible mixture in-troduced
into the intake passage 56 via the reed valve 64 is fed
into the intake passage 56 located downstream of the flow
control valve 57 via the small throttle valve 70, the
conduit 68, the inside chamber 67 of the accumulator 66
and the conduit 69. On the other hand, the burned gas
blowil~g back in-to the intake port 45 from the combustion
chamber 44 is fed into the inside chamber 67 of the accumu-
lator 66 via the conduit 69. Consequently, in the embodiment
illustrated in Fig. 13, the burned gas and the fresh
combustible mixture are mixed with each other and, thus,
the heat exchanging operation therebetween is carried out.
- 27 -
As a result of this, the vaporization of fuel is promoted
and, at the same time, the fresh combustible mixture is
reformed. In the embodiment illustrated in Fig. 13, it is
preferable that the volume of -the inside chamber 67 of -the
accumulator 66 be larger than that of the combustion
chamber 44 when the piston is positioned a-t the bottom
dead center. That is, by setting the volume of the inside
chamber 67 at the above-mentioned size, the unburned gas
and the frQsh combustible mixture can remain in the inside
chamber 67 of the accumulator 66 for a long time, whereby
the vaporization of fuel is further promo-ted and, at the
same time, the combustible mixture is fur-ther reformed.
Fig. 14 illustrates a further embodiment of a
4-cycle engine. In the embodiment illustrated in Fig. 14,
a conduit 69' is elongated as compared with the conduit 69
illustrated in Fig. 13, and an accumulator 66' has an
inside chamber 67' having a volume which is smaller than
the volume of the inside chamber 67 of the accumulator 66
illustrated in Fig. 13. As mentioned above, the conduit
69' in the embodiment illustrated in Fig. 14 has a consider-
ably long length and, thus, the burned gas and the fresh
combus-tible mixture are gradually fed into the intake port
45 while reciprocally moving in the conduit 69'. In this
embodiment, since the reciprocal movement of the unburned
gas and of the fresh combustible mixture is created in the
conduit 69', the vaporization of fuel is further promoted
and, at the same time, the com`bustible mixture is further
reformed as compared to the embodiment illustrated in
Fig. 13. On the other hand, in the 4-cycle engine illustrated
in Figs. 13 and 14, the flow control valve 57 remains
- 28 -
'
~s-
fully opened when the engine is operating under a heavy
load. Consequently, when the engine illustrated in Figs.
13 and 14 is operating under a heavy load, a high outpu-t
power similar to that in a conventional engine can be
obtained.
Fig. 16 illustrates the case wherein the present
invention is applied to a rotary piston engine. Referring
to Fig. 16, 80 designates a housing, 81 a rotor rota-ting
in the direction T-and having three corners which slide on
the inner wall of the housing 80, 82 three combustion
chambers, formed between the housing 80 and the rotor 81,
and 83 a pair of spark plugs; 84 designates an intake port
opening into the combustion chamber 82, 85 an intake
branch passage connected to the intake port 84, 86 an
intake manifold and 87 a carburetor; 88 designates a
throttle valve of the carbure-tor 87, 89 an exhaust port
opening into the combustion chamber 82, 90 an exhaust
passage connected to the exhaust port 89 and 91 an exhaust
manifold. An arm 92 is fixed onto the throttle valve 88
and the tip of the arm 92 is connected to the accelerator
pedal 79 via a wire 93. -A flow control valve 95 is disposed
in an intake passage 94 located downstream of and near the
throttle valve 88, and is fixed onto a valve shaft 96
pivotably mounted on the intake manifold 86. A cam 97 is
mounted on the valve shaft 96 and a wire 98, which is
wound on the outer periphery of the cam 97, is connected
to the accelerator pedal 79~ The relationship of -the
opening area ratios of the throttle valve 88 and the flow
control valve 95 is equal to the relationship between the
opening area ratios of the opening area ratio Y of the
~ 29 -
- ~ .
throttle valve and the opening area ratio X of the scavenging
control valve, shown by the curved line C in Fig. 6 and
aescribed with reference to Fig. 3. Consequently, the
flow control valve 95 is gradually opened and then, fully
opened before the throttle valve 88 reaches a point corre-
sponding to the opening area ratio of approximately 40
percent. On the other hand, the flow con-trol valve 95
remains fully opened when the opening area ratio of the
throttle valve 88 is larger than approximately 40 percent.
In a rotary piston engine illustrated in Fig. 16,
the exhaust-port 89 and the intake port 84 open into the
same combustion chamber 82 at the end of the exhaust
stroke. Consequently, when the level of the vacuum in the
intake manifold 86 is large, that is, when the engine is
operating under a partial load, a large amount of the
burned gas blows back into the in-take branch passage 85
from the combustion chamber 82 via the intake port 84 at
the end of the exhaust stroke. Since a large amount of
the burned gas blows back into the intake manifold 86, the
vaporization of the liquid fuel in the intake manifold 86
is-promoted and, at the same time, the fresh combustible
mixture in the intake manifold 86 is reformed. In addition,
the promotion of the vaporization causes a uni~orm distri-
bution of the fuel into the plurality of cylinders and
also causes an improvement in the responsiveness of the
engine with respect to the depressing opera-tion of -the
accelera-tor pedal 79. Further, the promotion of the
vaporization enables a stable combustion to be obtained
even if a lean ai~-fuel mixture is used. Particularly at
the time of idling and a-t the time when the engine is
- 30 -
6~
operating under a light load, wherein a sa-tisEactory
vaporizating operation of fuel cannot be obtained in a
conventional engine, the distribution oE fuel into the -
cylinders becomes uniform in an engine according -to the
5 present invention. As a result of this, since a good
combustion can be obtained in all of the cylinders, the
fuel consumption is grea-tly improved.
~ s mentioned above, in order to ef~ectively promote
the vaporization of fuel and reform the combustible mixture,
it is preferable that a large amount of the burned gas be
caused to blow back into the intake manifold 86. To this
end, it is preferable that an exhaust control valve 99 be
disposed in an exhaust pipe 91' as illustrated in Fig. 16.
The exhaust control valve 99 is caused to rotate by means
of a wire 100 connected to the accelerator pedal 79, in
the same manner as the exhaus-t control valve ?9 illustrated
in Fig. 8. The relationship between the opening area
ratios of the exhaust control valve 99 and the throttle
valve 88 is equal to the relationship between the opening
area ratio X of the exhaust control valve and the opening
area ratio Y of the throttle valve, shown by the curved
line D in Fig. 6 and described with reference to Fig. 3.
Consequently, the flow rate of the exhaust gas is prevented
by the exhaust control valve 99 when the exhaust control
valve 99 is slightly opened, tha-t is, when the engine is
operating under a light load, and the pressure in the
combustion chamber 82 at the end of the exhaust stroke is
greater as compared with the case wherein -the exhaust
control valve 99 is fully opened. As a result of this, a
large amount of the burned gas can be caused to blow back
- 31 -
67~5
into the intake manifold 86.
Fig. 17 illustrates another embodiment of a rotary
piston engine in which the vaporiza-tion of fuel can be
further promoted and the fresh combustible mixture can be
further reformed. Referring to Fig. 17, a reed valve 101
only permitting the downward flow of the fresh combustible
mixture is provided, and an accumulator 66 having a diaphragm
102 therein is provided. An inside chamber 104 of the
accumulator 103 is connected via a conduit 105 to the
intake passage 94 upstream of the flow control valve 95,
on one hand, while the inside chamber 104 is connected via
a conduit 106 to the intake passage 94 downstream of the
flow control valve 95, on the other hand. A small throttle
valve 107 is disposed in the conduit 105 and is fixed onto
a valve shaft 108 pivotably mounted on the conduit 105. A
cam 109 is mounted on the valve shaft 108 and a wire 110
connected to the accelerator pedal 79 is wound on the
outer periphery of the cam 109. The relationship between
the opening area ratios of the throttle valve 88, the flow
control valve 95, the small throttle valve 107 and the
exhaust control valve 99, is equal to the relationship
between the opening area ratios of the corresponding
throttle valve 49, flow controI valve 57, small throttle
valve 70 and exhaust control valve 61 illustrated in
Fig. 13, which relationship is shown in Fig. 15. In this
embodiment, the exhaust control valve 99 is arranged in
the exhaust manifold 91.
In this embodiment, in the same manner as described
with reference to Fig. 13, since the flow control valve 95
remains fully closed when the engine is operating under a
- 32 -
&5
partial load, the fresh combustible mixture introduced
into the intake passage 94 via the reed valve 101 is fed
into the intake passage 94 downstream of the flow control
valve 95 via the small throttle valve 107, the conduit
105, the inside chamber 104 of the accumulator 103 and the
conduit 106. On the other hand, the burned gas blowing
- back into the intake branch passage 85 from the combustion
chamber 82 at the end of the exhaust stroke is fed into
the inside chamber 104 of the accumulator 103 via the
conduit 106. Consequently, the unburned gas and the fresh
combustible mixture are mixed with each other and, thus,
the heat exchanging operation therebetween is carried out.
As a result of this, the vaporization of fuel is promoted
and, at the same time, the fresh combustible mixture is
reformed. In addition, as is described with reference to
Fig. 13, in the embodiment illustrated in Fig. 17, it is
also preferable that the volume of the inside chamber 104
of the accumulator 103 be larger than the total volume of
the combustion chambers 82. Furthermore, in the embodiment
illustrated in Fig. 17, the conduit 106 may be elongated
similar to the conduit 69' illustrated in Fig. 14.
In all of the embodiments hereinbefore described,
when the engine is operating under a heavy load, the
operation of the engine according to the present invention
is equal to that of a conventional engine. Consequently,
in all of the embodiments of the engine according to the
present invention, the engine may be provided with an
exhaust gas recirculating device for recirculating the
exhaus-t gas into the intake system of the engine only when
the engine is operating under a heavy load.
7~;S
The active -thermoatmosphere combustion causes the
reduction in the amount of harmful HC components in the
.. . . . ..
exhaust gas and also causes a considerable improvement in
the fuel consumption. In addition" even if a lean air-fuel
mixture is used, since an active thermoatmosphere combus-tion
is caused, the amount of harmful NOX components can be
reduced. Particularly in a multi-cylinder engine, since
the distribution of fuel into the cylinders becomes uniform,
a lean air-fuel mixture can be used as mentioned previously
and, at the same time, a stable combustion can be ob-tained
in all of the cylinders. As a result of this, the irregu-
larity in the torque generated in the respective cylinders
is extremely minimized and the vibration of the engine is
reduced. In addi-tion, when the active thermoatmosphere
combustion is carried out, the ignition delay does not
occur and, as a result, a quiet operation of the engine
can be effected at the time of idling and at the time when
the engine is operating under a partial load.
-While the invention has been described by reference
to specific embodiments chosen for purposes of illustration,
- it should be apparent that numerous modifications could be
made thereto by those skilled in the art without departing
from the spirit and scope of the invention.
- 34 -
.
