Direct injection internal combustion engine of compression ignition type

A direct injection internal combustion engine of a compression ignition type in which portions of an injected fuel spray which have not been evaporated are prevented from striking the walls of the combustion chamber, and the fuel spray is prevented from catching up the air around it when moving in the combustion chamber, thereby providing an engine having reduced combustion noise, reduced amount of smoke and noxious emissions, and improved fuel economy. A combustion chamber is formed in the piston substantially in the form of a spherical cavity gradually narrowing towards the opening of the cavity at the top surface of the piston. An air intake mechanism swirls intake air supplied to the combustion chamber. A swirl injector injects a fuel spray substantially in the form of a hollow cone having a velocity component in a direction tangential of the central axis of the nozzle. The nozzle is offset from the central axis of the cavity and inclined with respect to the central axis of the cavity so that the fuel spray is injected in a forward direction of a swirl stream of intake air in the cavity.

BACKGROUND OF THE INVENTION 
The present invention relates to a direct injection internal combustion 
engine of a compression ignition type in which fuel is injected into a 
cavity formed in a piston to form an excellent mixture gas. 
A high-speed direct injection diesel engine of relatively small size is 
known which has a toroidal combusion chamber in the form of a 
semi-cylindrical cavity in the top of a piston. A conical protrusion is 
formed at the center of the bottom of the cavity to provide an annular 
recess. A multi-hole nozzle (having four or five holes) is arranged, as a 
fuel injection valve, at the center of the cavity. The nozzle injects a 
spray of fuel radially, and the fuel evaporates while passing across a 
swirl stream formed in the cavity, thus forming mixture gas. 
The fuel spray injected by the nozzle has a large velocity and a strong 
penetration force in the axial direction of the nozzle hole. Therefore, 
while passing through the high temperature air in the cavity, the fuel 
spray evaporates gradually, the diameters of the fuel droplets in the 
spray decrease, and the flight velocity of the droplets also decreases. 
Accordingly, the fuel spray is caused to flow by the swirl stream formed 
in the cavity, thus forming mixture gas in the swirl stream near the 
downstream walls of the cavity. 
Immediately after this conventional diesel engine is started, namely when 
the engine is operating at low temperature, the temperature of the walls 
of the cavity is also low, the air temperature is not elevated during the 
compression stroke, and thus the temperature of the swirl stream formed in 
the cavity is not so high. Accordingly, the speed of evaporation of the 
sprayed fuel droplets is low. In other words, the fuel spray does not 
sufficiently evaporate, as a result of which large unevaporated fuel 
droplets strike against the walls of the cavity. Accordingly, smoke is 
generated, the rate of pressure rise is high, the noise level is high, and 
undesired hydrocarbons (HC) are formed. These are disadvantages 
accompanying the conventional diesel engine. 
Also, for normal low-speed operation, the swirl stream formed in the cavity 
is weak, and in this case too the sprayed fuel droplets strike against the 
walls of the cavity causing the same difficulties as for cold running. 
The swirl stream in the cavity is generally formed by an intake port which 
is helical in configuration, namely, a helical port. However, during low 
speed operation of the engine in which the piston is moving at a low 
speed, the velocity of the swirl stream is low, and therefore the swirl 
stream cannot sufficiently decelerate the radial movement of the fuel 
spray imparted by the multi-hole nozzle so that in this case as well the 
sprayed fuel droplets strike against the walls of the cavity. 
On the other hand, during high speed operation of the engine, the piston is 
moving at a high speed, and therefore the velocity of the swirl stream 
formed in the cavity is high. In this case, mixture gas layers formed near 
the walls of the cavity are caused to flow; that is, the mixture gas 
layers formed by adjacent injection holes tend to overlap, forming a 
mixture gas of excessively high concentration, with the result that smoke 
is generated. 
In small engines for automobiles, and especially in engines whose cylinder 
diameters are of the order of 75 to 100 mm, the range of engine running 
speed is wide; the idling speed is 500 to 800 rpm and the maximum speed is 
4000 to 5000 rpm. Thus, such engines have a problem of forming a mixture 
gas of excessively high concentration. 
The fuel spray injected by only one of the injection holes of the 
multi-hole nozzle will now be considered. The sprayed fuel droplets have a 
generally very high velocity, which causes air to be pushed in front of 
the spray. The velocity of the fuel spray decreases as the fuel spray 
evaporates. As a result, fuel spray is caught up in and driven by the 
swirl stream, thus forming mixture gas. Ignition is initiated near the 
high temperature wall of the cavity near the side surface of the front end 
of the spray. There is a sufficient amount of air on the surface of the 
fuel spray, which is in the form of a solid cone, and therefore after 
ignition near the front end of the fuel spray, not only the mixture gas 
layer located downstream of the swirl stream but also the surface layer of 
the fuel spray which has been just injected is combusted. That is, 
immediately after ignition the velocity of combustion is so high that 
there is no time to efficiently use the available intake air in the 
cavity, and accordingly smoke is liable to be generated. Furthermore, as 
the rate of pressure rise is high, the noise of combustion is also large. 
In another type of conventional direct injection diesel engine, a 
combustion chamber is provided by forming a three-quarter spherical cavity 
in the top of the piston. The air is swirled by a helical port so that a 
swirl stream is formed in the cavity at the end of the compression stroke. 
This diesel engine has found practical use in some fields. 
In this diesel engine, the fuel, injected by a single-hole nozzle or a 
double-hole nozzle, is sprayed against the walls of the cavity, thus 
forming a fuel film on the wall. The fuel film is evaporated on the cavity 
wall. The vapor of the fuel thus evaporated is driven by the swirl stream 
inside the cavity to form a mixture gas. 
Compared with the first-described diesel engine having a toroidal 
combustion chamber, the second-described diesel engine is advantageous in 
that, as only the fuel vapor which is formed by evaporation of the fuel 
film is mixed with air for combustion, the quantity of smoke generated is 
small. However, the latter diesel engine is still disadvantageous in that 
the combustion characteristics are strongly affected by the temperature of 
the wall, the engine is difficult to start at low temperatures, and a 
large amount of noxious emissions, particularly hydrocarbons, is exhausted 
immediately after starting. 
In order to solve the above-described problems accompanying conventional 
small compression ignition type direct injection internal combustion 
engines, the inventors have conducted systematic experiments, analyses and 
trial manufacture, and accomplished this invention. 
SUMMARY OF THE INVENTION 
An object of the present invention is to eliminate the above-described 
difficulties accompanying a conventional diesel engine, that is, to 
prevent the portions of the fuel spray which are not evaporated from 
striking the walls of the combustion chamber to prevent as much as 
possible the fuel spray from catching up the air around it while the fuel 
spray is moving in the combustion chamber, to reduce the rate of 
combustion pressure rise immediately after ignition, to decrease the 
combustion noise, and to reduce the quantity of smoke generated. 
Provided according to the invention is a direct injection internal 
combustion engine of a compression ignition type in which air is sucked 
into a combustion chamber and compressed by a piston, and fuel is injected 
thereinto so as to be ignited and combusted, which, according to the 
invention, comprises: an air intake mechanism having swirling means which 
swirls intake air supplied to the combustion chamber; a combustion chamber 
formed as a substantially spherical cavity in the top of a piston with the 
cavity gradually narrowing towards the opening thereof; and a swirl 
injector for injecting sprayed fuel droplets substantially in the form of 
a hollow cone which has a velocity component in the tangential direction 
from the injection port thereof which is offset from the central axis of 
the cavity and inclined with respect to the central axis of the cavity so 
that the fuel spray is injected in the forward direction of a swirl stream 
of intake air inside the cavity.

DESCRIPTION OF THE PREFERRED EMBODIMENTS 
In the direct injection internal combustion engine of the compression 
ignition type according to the invention, as shown in FIG. 1A, the air 
intake forms a swirl stream SW of intake air in the cavity C in advance. 
The nozzle of a swirl injector SN, which is positioned in the forward 
direction of the swirl stream and offset from the central axis of the 
cavity and inclined with respect to the central axis of the cavity, 
injects a hollow cone-shaped fuel spray which has a velocity component in 
the tangential direction, and which does not catch up the air around the 
fuel spray. This fuel spray has a small penetration force, low velocity of 
movement, and small droplet diameter, and is smoothly driven by the swirl 
stream without disturbing the flow pattern so that it is dispersed 
circumferentially of the cavity C in the form of a doughnut, as shown in 
FIGS. 1B and 2. The sprayed fuel droplets thus dispersed are further 
dispersed in the direction of the axis of the cavity (downwardly) by a 
squish flow S formed as the piston rises, and it flows into the cavity C 
through the opening from the flat part of the piston near the end of the 
compression stroke. While being dispersed circumferentially and downwardly 
of the cavity C, the fuel spray, which is a group of tiny fuel droplets, 
evaporates to mix with the intake air, forming a mixture gas layer in the 
form of a doughnut in the cavity, as shown in FIG. 1B. Ignition initiates 
at a point or a plurality of adjacent points near the front end of the 
mixture gas layer, and the fuel spray is injected in the direction of the 
swirl stream at the part of the cavity where the velocity of the swirl 
stream is higher. Since the velocity of the fuel spray itself is low, the 
fuel spray does not catch up the air around it. Therefore, during the 
initial period of combustion, the fuel spray which has been just injected 
and is in the form of a hollow cone will not be combusted immediately, 
that is, combustion gradually advances from the front end of the fuel 
spray, which has been dispersed in the form of a doughnut or 
circumferentially by the swirl stream, towards the upstream part as shown 
in FIG. 1B. Accordingly, the amount of smoke produced is small, and the 
rate of combustion pressure rise is also small, and therefore the noise 
level is low. In addition, as the air in the cavity is effectively 
utilized, the specific fuel consumption is reduced. These are some of the 
effects or merits of the internal combustion engine of the invention. 
Furthermore, in the internal combustion engine of the invention, the cavity 
C is substantially spherical. Therefore, the swirl stream SW and the 
squish flow S are formed uniformly along the wall of the cavity C to the 
extent that there is no dead space in the cavity C, which makes it 
possible to form a good mixture gas layer or region in the cavity C. 
The effects of the direct injection internal combustion engine of the 
compression ignition type will now be described in more detail. 
In the internal combustion engine of the invention, after the intake 
(suction) stroke, the swirl stream SW of intake air is formed in the 
cavity C, and after the start of the compression stroke, the squish flow S 
flows into the cavity C from the flat part of the top of the piston. 
The swirl injector injects a fuel spray, which has a small penetration 
force and a small velocity, into the cavity in the flow direction of the 
swirl stream formed, shown in FIG. 1A. As the fuel, having the swirling 
motion imparted by the slit of the injection valve and injected in the 
form of a hollow conical film parts from the nozzle, its diameter 
increases while its thickness decreases, as a result of which the film is 
broken into tiny droplets, that is, fuel spray is formed. The fuel spray 
injected by the swirl injector as described above have both a velocity 
component in the axial direction of the nozzle and a velocity component in 
the tangential direction. This fuel spray has a small velocity and 
penetration force, and accordingly a very small travel distance and good 
atomization compared with fuel spray injected by the conventional 
multi-hole nozzle. 
The sprayed fuel droplets injected in the direction of the swirl stream in 
the cavity, having a low velocity and a small penetration force, will not 
catch up the air around it. Therefore, when the fuel spray enters the 
swirl stream, as shown in FIGS. 1B and 2, it readily flows 
circumferentially in the cavity near the wall of the cavity and is 
smoothly driven by the swirl stream, thus being dispersed 
circumferentially in the cavity. 
The sprayed fuel droplets (atomized sprayed fuel droplets) are driven by 
the swirl stream in the cavity and are formed mainly in the middle and 
upper region. The sprayed fuel droplets thus dispersed form a mixture gas 
layer in the form of a doughnut along the walls of the combustion chamber 
while being moved by the swirl stream. 
Thus, the internal combustion engine of the invention is fundamentally 
different in its mixture gas forming mechanism from the conventional 
toroidal type engine in which the fuel spray forms the mixture gas while 
catching up the air around it with a strong penetration force, or the 
conventional engine in which a fuel film is formed on the walls of the 
cavity and then evaporated. 
As the fuel injected in the form of a hollow conical film by the swirl 
injector is atomized, its velocity quickly decreases and the fuel spray is 
reduced in volume because of the density of the gas surrounding it. As a 
result, the fuel spray, while mixing with the air and evaporating, is 
dispersed circumferentially in the cavity by the swirl stream so as to be 
distributed around the periphery of the cavity by the squish flow which 
flows in the same direction. 
Ignition initiates at one point or a plurality of adjacent points near the 
front end of the mixture gas. In the engine of the invention, unlike the 
conventional toroidal type engine in which fuel sprays are injected 
radially from the center of the cavity where the velocity of the swirl is 
low, the fuel spray is injected in the direction of the swirl stream to 
the peripheral part of the cavity where the velocity of the swirl stream 
is high and the velocity of the fuel spray is low, and therefore the fuel 
spray will not catch up the air around it. Accordingly, during the initial 
period of combustion, the fuel spray which has just been injected by the 
swirl injection will not burst into flame immediately; that is, the flame 
advances gradually from the ignition point towards the downstream end of 
the region where the mixture gas is formed through the dispersion of the 
fuel spray by the swirl stream. Thus, combustion gradually advances while 
effectively utilizing the intake air flow in the cavity. 
Accordingly, the internal combustion engine of the invention is 
advantageous in the following points: The quantity of smoke generated and 
exhausted is small, and the rate of combustion pressure rise is low, and 
therefore the amount of noise produced is small. In the cavity, the intake 
air and the sprayed fuel droplets are well diffused and mixed, and the 
intake air in the cavity is effectively used. Therefore, with this 
internal combustion engine, even a small quantity of fuel can be ignited 
for combustion. Furthermore, the output characteristics and the fuel 
consumption characteristics are improved with the use of the invention. 
The invention can be practiced as follows: 
In a direct injection internal combustion engine of a compression ignition 
type constructed according to a first embodiment of the invention, the 
intake valve arrangement as shown in FIGS. 3A and 3B is implemented with 
an offset port designed so that the side wall IS of the intake passage, 
which is near the center of the piston, gradually inclines outwardly, with 
an extension line of the inclined side wall extending within the region 
from the outer wall of the rod IR of an intake valve IV to a position 
apart from the outer wall thereof. The side wall OS of the intake passage, 
which is on the side of the periphery of the piston, is linear and 
smoothly merges with the arcuate side wall CS, the center of which is 
shifted toward the periphery of the piston with respect to the central 
axis of the intake valve IV. 
In the case of a conventional helical port, a large capacity spiral chamber 
is formed in the valve chamber and the intake passage is bent spirally. 
Therefore, the air resistance is high during high speed operation of the 
engine. However, as described above, in accordance with this first aspect 
of the invention, the intake valve arrangement includes an offset port. 
That is, no large-capacity spiral chamber is formed in the valve chamber, 
and therefore the sucked air flows smoothly without striking the rod IR of 
the intake valve IV. Thus, this aspect of the invention is advantageous in 
that the air resistance is low and the intake air volume efficiency is not 
lowered during high speed operation of the engine. 
In accordance with the first aspect of the invention, as shown in FIG. 3A, 
the intake port extends tangential to the cylinder. Therefore, the first 
aspect brings about the advantage in that the speed distribution of a 
swirl stream formed in the cylinder approaches that of a solid vortex, 
compared with a helical port, and the part of the swirl stream which moves 
at the highest speed closely approaches the inner wall of the cylinder. 
When the maximum speed distribution of the swirl stream is near the inner 
wall of the cylinder, the speed of the swirl stream in the cavity at the 
end of the compression stroke is increased. The air resistance of the 
intake mechanism depends on the maximum air velocity. Therefore, generally 
as the maximum speed distribution of the swirl stream closely approaches 
the periphery in the suction stroke, the intake air volume efficiency of 
the engine increases with the speed of the swirl stream in the combustion 
chamber unchanged. This brings about the practical advantage that the 
torque of the engine is increased during high speed operation of the 
engine. 
As described above, in accordance with the first aspect of the invention, 
scattering of the fuel spray circumferentially in the cavity depends on 
the swirl stream of intake air. Therefore, the engine requires a strong 
swirl stream compared with a conventional toroidal type engine. In 
general, the swirl ratio (obtained by dividing the angular velocity of 
swirl by that of the engine) used for measuring the strength of a swirl 
stream in the combustion engine should be in a range of 2.6 to 4, more 
preferably, 3 to 3.7. 
In a compression ignition type direct injection internal combustion engine 
constructed in accordance with a second aspect of the invention, the 
spherical cavity is designed so that, as shown in FIG. 4A, the 
configuration of its walls conforms to that of the shape of the sprayed 
fuel droplets injected by the swirl injector. 
In accordance with the second aspect of the invention, the wall of the 
spherical cavity conforms to the shape of the fuel spray injected by the 
swirl type injection nozzle as described above. Therefore, the second 
aspect of the invention brings about the advantage that the fuel spray 
will not stick to or collide with the walls of the cavity. The diameter of 
the opening CO of the cavity is smaller than the diameter of the cavity 
itself, and the opening has a cut EG as shown in FIG. 4B to prevent 
mechanical contact with the swirl type injection nozzle and blocking of 
the fuel spray. 
In a direct injection internal combustion engine of a compression ignition 
type constructed in accordance with a third aspect of the invention is 
that the position and direction of the swirl injector is determined so 
that, when fuel spray is formed by the swirl injector, the spray thus 
formed is confined between the central axis CC of the cavity and the wall 
CW of the cavity as shown in FIG. 4B. 
In accordance with the third aspect of the invention, dispersion of the 
fuel spray from the central axis CC of the cavity to the wall CW is 
achieved by appropriately setting the fuel spray angle from the swirl type 
injection nozzle SN, thus preventing the fuel spray from sticking to or 
striking against the wall CW of the cavity and preventing the spray from 
passing beyond the central axis CC of the cavity. 
In accordance with a fourth aspect of the direct injection internal 
combustion engine of the compression ignition type of the invention, the 
opening of the cavity is formed so that the cavity contraction ratio (the 
ratio of the arena of the opening of the spherical cavity to the total 
area of the top of the piston) is in a range of 0.05 to 0.12. 
In accordance with the fourth aspect of the invention, the cavity's 
contraction ratio is selected to be in the above-described range. This is 
advantageous in that, with the aid of the squish flow, the fuel spray is 
supplied towards the bottom of the cavity of compensate for the force of 
penetration of fuel injected by the swirl type injection nozzle, which is 
relatively reduced compared with the strength of the swirl stream for high 
speed operation of the engine. If the squish flow is excessively strong, 
then the strength of the turbulent flow in the cavity is increased, as a 
result of which the thermal loss is increased. The fuel consumption is 
also increased. 
In accordance with a fifth aspect of a direct injection internal combustion 
engine of a compression ignition type of the invention, the cavity is 
formed so that the ratio of the volume of the cavity to the total 
clearance volume between the piston at the top dead center and the 
cylinder head is in the range of 0.6 to 0.80. 
Accordingly, with the volume of the cavity determined as described above, 
the quantity of intake air necessary for combustion is held in cavity, 
thus preventing any decrease of torque and eliminating difficulties 
involved in the manufacture of the engine and collision of the piston 
against the cylinder head when the engine is running. 
In accordance with the a sixth aspect of a direct injection internal 
combustion engine of a compression ignition type of the invention, as 
shown in FIG. 5, the central axis of the nozzle of the swirl type 
injection nozzle SN is extended inside the wall of the opening CO of the 
cavity and outside a circle whose diameter is 0.8D, where D is the 
diameter of the opening CO. 
As is apparent from the above description, in accordance with the sixth 
aspect of the invention, the nozzle of the swirl type injection nozzle is 
directed to the region in the cavity where the swirl stream is strong. 
This brings about the advantage that the squish and swirl streams disperse 
the sprayed fuel droplets circumferentially and downwardly in the cavity, 
thus forming excellent mixture gas, specifically, the sprayed fuel 
droplets are dispersed circumferentially and the strong squish flow is 
utilized near the top dead center. 
In accordance with a seventh aspect of a direct injection internal 
combustion engine of a compression ignition type of the invention, as 
shown in FIG. 6, the angle .theta. between the central axis SO of the 
nozzle of the swirl injector and the central axis CO of the cavity is in a 
range of 30.degree. to 70.degree.. 
With the angle .theta. of the nozzle of the swirl injector with respect to 
the central axis of the cavity defined as described above, the seventh 
aspect is advantageous in that the fuel spray will not stick to or strike 
against the wall CW of the cavity C. 
A first embodiment of the above-described compression ignition type direct 
injection internal combustion engine will be described. 
A specific feature of this embodiment, represented by the first through 
seventh aspects of the invention enumerated above, resides in that a swirl 
stream is formed in advance in a cavity 2 in the top of a piston 1 so as 
to cause a fuel spray injected by a swirl type injection nozzle 3 to 
disperse circumferentially in the cavity 2, with the aid of a squish flow, 
which is maintained until the end of the compression stroke. The fuel 
spray is dispersed downwardly in the cavity 2 so that the fuel is 
sufficiently evaporated to mix well with the intake air, thereby forming 
excellent mixture gas in the cavity 2. The mixture gas thus formed is 
ignited to gradually burn the fuel spray dispersed downstream thereof. 
These actions will be described in more detail with reference to FIGS. 7, 8 
and 9. As shown in FIG. 7, the substantially spherical cavity 2 is formed 
in the flat top of the piston 1, which reciprocates in the cylinder 5, 
substantially at the center of the piston. The cavity 2 is shaped such 
that, as shown in FIG. 7, its cross section is in the form of a circle 
whose diameter varies with the level of the cross section, and the 
vertical section has the configuration of two ellipses of different 
curvatures on opposite sides of the maximum diameter 2m. The diameter of 
the opening 2h of the cavity 2 is smaller than those of other parts, and 
the total area of the opening, including a cut described below, is 9.8% of 
the total area of the piston 1. This is done to make the squish flow S 
flowing into the cavity 2 more suitable. In order to prevent the swirl 
type injection nozzle 3 and the fuel spray from the outlet of the 
injection nozzle 3 from striking against the opening 2h, a substantially 
U-shaped cut 2C is formed in the top of the piston merging with the 
opening 2h near the nozzle of the swirl injector as shown in FIG. 8A. 
Furthermore, by taking the utilization of air in combustion into account, 
the dimensions of the cavity 2 are selected so that the volume thereof is 
72% of the volume of the combustion chamber with the piston at the top 
dead center. 
In the above-described engine, the compression ratio, determined by the 
volumes when the piston is at top dead center and at bottom dead center, 
is set to 17. In this engine, due to the employment of the swirl type 
injection valve, fuel is atomized sufficiently, and thus it is unnecessary 
to set the compression ratio to a high value. If the compression ratio is 
excessively high, then the ignition delay time becomes short, thus 
resulting in the generation of smoke. If the compression ratio is 
excessively low, ignition is not carried out satisfactorily. Thus, the 
compression ratio should be in the range of 14 to 20. 
The fuel injector is a slit swirl type injection nozzle 3 which, as shown 
in FIG. 7, penetrates the cylinder head 4 and is mounted with its outlet 
confronting the cut 2c of the cavity 2. That is, the opening of the swirl 
type injection nozzle 3 extends in the direction of forward flow of the 
swirl stream SW formed in the cavity 2, and the tip of the nozzle is 
offset by 0.18D (D is the diameter of the piston) from the central axis of 
the cavity. The length of a perpendicular line from the central axis of 
the nozzle to the central axis of the cavity 2 is 0.12D. The angle .theta. 
between the central axis of the injection port and the central axis of the 
cavity 2 is 45.degree.. 
The swirl type injection nozzle 3, as shown in FIG. 9, is composed of a 
nozzle body 30, which is a hollow cylindrical member which is elongated at 
the end, and a needle member 31, which is a stepped rod inserted into the 
nozzle body 30. A swirl chamber 35 is formed in the end portion of the 
nozzle body 30, and an injection hole 32 is formed in the lower end of the 
nozzle body 30 coaxial with the nozzle body and communicating with the 
swirl chamber 35. The needle member 31 has a conical end 33 which abuts 
against the injection hole 32 in such a manner as to close the latter. The 
conical end 33 is formed by machining the larger-diameter part of the 
needle member 31. Two slits 34 (only one of which is shown in FIG. 9) are 
cut in the cylindrical wall of the larger diameter part of the needle 
member 31 in such a manner that they form a predetermined angle with 
respect to the central axis of the needle valve so that the swirl chamber 
35 is communicated with a chamber 36, which is in turn communicated 
through a fuel supply passage 37 to a distribution type fuel pump (not 
shown). 
In the swirl injector, the angle, sectional area and length of the slits, 
the dimensions of the swirl chamber 35 and the diameter and length of the 
injection hole 32 are determined so that, with the distance between the 
center of the cavity 2 and the wall taken into account, the injection 
valve forms a hollow conical spray pattern, and the spray angle .alpha. 
(FIG. 9) of the spray pattern is 40.degree. (at atmospheric pressure). It 
has been found through experiments that the best result is obtained when 
the diameter of the injection hole 32 is in the range of 0.3 mm to 1.0 mm. 
In the present example, the diameter of the injection hole 32 is 0.6 mm. 
The spray angle .alpha. of the fuel is effective when it is in the range of 
25.degree. to 60.degree. (at atmospheric pressure). If the angle is 
excessively small, the fuel is not sufficiently dispersed 
circumferentially of the cavity 2, and if it is excessively large, the 
density of the mixture gas at the center of the cavity is increased, which 
may lead to the phenomenon of so-called thermal pinch. 
In FIG. 9, an angle .beta. represents the thickness of the hollow conical 
fuel spray. The angle .beta. should not be large for it affects the 
atomization, and preferably is in the range of 5.degree. to 15.degree.. 
The total injection period of the swirl type injection valve is set to 
about 12.degree. for an engine speed of 1200 rpm, and to about 29.degree. 
for an engine speed of 4000 rpm. The injection pressure (pressure at the 
valve opening) is about 160 kg/cm.sup.2. 
As shown in FIG. 7, the intake valve 6 is inserted into the cylinder head 
4, and an offset port 7 having a predetermined swirl ratio 3.6 is formed 
in the intake passage where the intake valve 6 is arranged. 
The offset port 7 is similar to that shown in FIGS. 3A and 3B, the former 
differing from the latter only in the direction of arrangement. Therefore, 
a detailed description of the offset port 7 will not be made. 
In the direct injection internal combustion engine of the compression 
ignition type thus constructed, the swirl stream SW is formed in the 
cavity 2 by the offset port 7 in advance, and a substantially hollow 
conical fuel spray having velocity components in tangential directions and 
a suitable spray angle is injected from the injection hole 32 of the swirl 
injector 3 in the forward direction of the swirl stream SW so that, as 
shown in FIG. 8B, the swirl stream SW smoothly disperses the sprayed fuel 
droplets circumferentially of the cavity 2 in such a manner that the 
sprayed fuel droplets is dispersed substantially in the form of a ring. 
The squish flow S which flows into the cavity through the opening 2h from 
the flat part of the top of the piston 1 near the end of the compression 
stroke further disperses the sprayed fuel droplets, which has been 
dispersed circumferentially of the cavity 2, downwardly in the cavity 2. 
That is, in the example, the fuel is dispersed as follows: The swirl type 
injection nozzle is utilized to disperse the sprayed fuel droplets 
radially of the cavity, the swirl stream SW is utilized to disperse the 
sprayed fuel droplets circumferentially of the cavity, and the squish flow 
S is used to disperse the sprayed fuel droplets downwardly of the cavity. 
During this dispersion, the sprayed fuel droplets evaporate gradually. 
However, it should be noted that the engine is designed so as to prevent 
the sprayed fuel droplets from sticking to or striking against the wall of 
the cavity 2 and to prevent overlapping of the sprayed fuel droplets, 
thereby preventing the formation of an excessively high-density mixture 
gas. In the example, the mixture gas in the cavity 2 is formed into an 
annular shape. 
Approximately 0.7 to 1 msec after the injection of fuel, ignition occurs at 
the front end of the fuel spray. The velocity of the swirl stream SW 
formed in the cavity 2 is high, and the velocity and the penetration force 
of the hollow conical sprayed fuel droplets in the axial direction of the 
nozzle are small, and therefore the hollow conically sprayed fuel droplets 
do not significantly catch up the air around them. Accordingly, because of 
this mixture formation, unlike the conventional engine using a toroidal 
system, the hollow conically sprayed fuel droplets which have just been 
injected by the swirl injector and have not evaporated yet are not 
surrounded immediately by the flame. That is, the combustion advances in 
the forward direction of the swirl and then gradually in a direction 
opposite the direction of the swirl toward the end of the combustion 
period, burning a part of the sprayed fuel droplets which are forming 
mixture gas while being dispersed circumferentially of the cavity 2 by the 
swirl stream SW. Accordingly, in the engine of the invention, unlike the 
conventional engine, the amount of smoke and soot formed is very small. 
If the spray angle of the spray of the swirl type injection valve is 
excessively large, then the spray will reach the central region of the 
cavity 2. The intake air does not reach this region owing to the 
centrifugal force of swirl rotation so that the spray does not mix with 
the air sufficiently, and therefore soot is formed therein. In order to 
eliminate this difficulty, in the example under discussion, the spray 
angle is set to 40.degree.. In general, the output of a diesel engine is 
limited by the concentration of soot. Therefore, the output can be 
increased by suppressing the formation of soot; that is, a high power 
diesel engine can be obtained. 
In the above-described engine, in the range of from low speed to high 
speed, combustion is carried out by efficiently using the air in the 
cavity 2. Therefore, the specific fuel consumption of the engine is 
decreased, and the output torque increased. Furthermore, in the engine of 
the invention, unlike the conventional engine in which combustion occurs 
explosively, the combustion advances gradually. Therefore, the engine of 
the invention is advantageous in that it has a low rate of pressure rise 
and level of combustion noise. Furthermore, in the above-described engine, 
the overall noise level generated in the idling state, which if high is 
very disturbing to the operator, is lower by as much as 3 to 4 dB than in 
an ordinary diesel engine, and especially the combustion noise in the 
frequency band of 1 to 5 kHz is extremely low. 
While preferred embodiments of the invention have been described in detail, 
it is particularly understood that the invention is not limited thereto or 
thereby, and it is evident to those skilled in the art that various 
changes and modifications may be made therein without departing from the 
invention.