Engine cooling system and construction machine

First cooling air (50) enters an engine room (1) through a cooling air inlet port (7a) from the exterior of the engine room (1), and is throttled by a suction tube (8a) after passing heat exchangers such as an intercooler (6a), an oil cooler (6b) and a radiator (6c), followed by entering a centrifugal fan (4). The first cooling air (50) is then blown off toward an outer circumference of the centrifugal fan (4). Second cooling air enters the engine room (1) through cooling air inlet ports (7b, 7b) from the exterior of the engine room (1), and flows around an engine (5), various electric equipments such as an alternator (10), and an oil pan (14) while cooling them. The second cooling air is then throttled by a suction tube (8b) before entering the centrifugal fan (4), and blown off toward the outer circumference of the centrifugal fan (4). Two streams of the first and second cooling air (50, 51) are discharged to the exterior through an exhaust port (9). This arrangement enhances the degree of sealing on the engine (5) side to reduce noise, and improves the effect of cooling the electric equipments (10) for higher reliability of the electric equipments (10).

TECHNICAL FILED 
The present invention relates to an engine cooling system, particularly a 
cooling system for engines equipped on automobiles or construction 
machines, for example, and a construction machine employing the engine 
cooling system. 
BACKGROUND ART 
Known techniques relating to the above type of engine cooling system are, 
for example, as follows. 
(1) JP-A-5-288053 
This known technique discloses that, in an engine cooling portion of a 
construction machine, cooling air is supplied to a heat exchanger by an 
axial fan coupled to an engine crankshaft through a fan belt. 
(2) JP-A-5-248239 
This known technique discloses that, in an engine cooling portion of a 
working vehicle such as an agricultural tractor, a centrifugal fan is used 
as a fan for supplying cooling air to enhance a cooling ability. 
(3) JP-A-5-248242 
With this known technique, a single centrifugal fan is employed such that 
cooling air taken in from the front of an agricultural machine and having 
passed a radiator is sucked into the centrifugal fan from the front 
thereof, and cooling air from an engine room rearward of the centrifugal 
fan is sucked into the centrifugal fan from the back thereof. 
DISCLOSURE OF THE INVENTION 
The above known techniques have problems below. 
Generally, in an engine room, an air cooling ability required for cooling 
external surfaces of an engine is much smaller, e.g., about 1/3 or less, 
than an air cooling ability required for cooling heat exchangers such as a 
radiator and an oil cooler. In the above known techniques (1) and (2), 
however, because cooling air heated to high temperature after cooling the 
heat exchangers such as the radiator flows into the engine side as it is, 
the flow rate of cooling air is the same between the heat exchanger side 
and the engine side. Stated otherwise, because the cooling air is 
discharged at a large flow rate necessary for cooling the heat exchanger 
side in its entirety through an exhaust port formed on the engine side, an 
opening defining the exhaust port has a very large size. This results in a 
reduction in the degree of sealing on the side of the engine that is the 
greatest noise source, and hence a difficulty in reducing noise. 
Also, many electric equipments, etc. are usually mounted in the engine room 
around the engine. In the above known techniques (1) and (2), however, 
because air heated to high temperature after cooling the heat exchangers 
such as the radiator flows around the electric equipments, the effect of 
cooling the electric equipments is so reduced as to raise a problem in 
reliability of the electric equipments. 
On the other hand, such a problem is avoided in the above known technique 
(3) because cooling air for cooling the electric equipments around the 
engine is introduced through a different flow path from that of cooling 
air passing the radiator and high-temperature air heated after cooling the 
radiator is not introduced to the electric equipments. However, the above 
known technique (3) has another problem as follows. 
The centrifugal fan employed in the above known technique (3) is 
illustrated as functioning as a double-suction centrifugal fan which sucks 
cooling air from two directions, i.e., from the radiator side and the 
engine side. In substance, however, the known centrifugal fan is made up 
of one single-suction centrifugal fan having one impeller. This impeller 
is designed to positively suck air on the radiator side and induce cooling 
air. For air on the engine side, however, a gap is defined on the back 
side of the fan, allowing air to be merely introduced through the gap 
under negative pressure. Thus, the impeller is not structured so as to 
positively suck air on the engine side and induce cooling air. As a result 
of such a structure, the flow rate of cooling air on the engine side is 
much smaller than that of cooling air on the radiator side. Also, the 
amounts of cooling air sucked from both the directions are instable in 
relative relation. Accordingly, the electric equipments, etc. cannot be 
cooled sufficiently. 
The present invention intends to solve the foregoing problems encountered 
in the known techniques, and its object is to provide an engine cooling 
system and a construction machine, which can improve the degree of sealing 
on the engine side for a reduction in noise, and can sufficiently cool 
electric equipments, etc. 
To achieve the above object, according to the present invention, there is 
provided an engine cooling system comprising at least one heat exchanger 
disposed in an engine room with an engine mounted therein and including a 
radiator for cooling water to cool the engine, and a cooling fan rotated 
through a rotary shaft for inducing cooling air to cool the heat 
exchanger, the engine, and electric equipments disposed around the engine, 
wherein the cooling fan is a centrifugal fan of a centrifugal 
double-impeller structure for inducing first cooling air taken into the 
engine room through a first air inlet port to cool the heat exchanger, and 
second cooling air taken into the engine room through a second air inlet 
port to cool the engine and the electric equipments, and the centrifugal 
fan is constructed to meet W1:W2=2:0.5-2:1.5 on an assumption that flow 
rates of the first cooling air and the second cooling air induced by the 
centrifugal fan are W1 and W2, respectively. 
Specifically, the first cooling air induced by one impeller of the 
centrifugal double-impeller structure is taken into the engine room 
through the first air inlet port for cooling the heat exchanger. After 
that, the first cooling air is sucked into the first impeller from one end 
side of the rotary shaft and then blown off toward an outer circumference 
of the fan. Also, the second cooling air induced by the other impeller of 
the centrifugal double-impeller structure is taken into the engine room 
through the second air inlet port for cooling the engine and the electric 
equipments, etc. around the engine. After that, the second cooling air is 
sucked into the second impeller from the other end side of the rotary 
shaft and then blown off toward the outer circumference of the fan. 
With the above arrangement, a cooling flow path of the second cooling air 
for cooling the engine side can be separated from a cooling flow path of 
the first cooling air for cooling the heat exchanger. Therefore, it is no 
longer required to discharge cooling air at a large flow rate from the 
engine side unlike conventional, and the flow rate of the second cooling 
air can be set to a smaller value just required for cooling the engine 
side. Accordingly, the size of an opening to be formed on the engine side 
can be set smaller than conventional, which results in the improved degree 
of sealing on the engine side and a reduction in noise. 
Also, since the cooling flow path of the second cooling air for cooling the 
engine side has a different route from the cooling flow path of the first 
cooling air for cooling the heat exchanger, high-temperature air heated 
after cooling the heat exchanger is prevented from flowing around the 
electric equipments, etc. unlike conventional. It is thus possible to 
enhance the effect of cooling the electric equipments, etc. and to improve 
the reliability of the electric equipments, etc. 
It is generally known that, of the total heat value 100% generated by an 
engine, the percentage of the heat value consumed by cooling of radiator 
water is about 30% and the percentage of the heat value consumed as 
radiation heat from outer walls of the engine is about 15%. Here, the 
former percentage practically corresponds to the heat value cooled by the 
first cooling air in the radiator and the latter percentage practically 
corresponds to the heat value cooled by the second cooling air around the 
engine. Therefore, on an assumption that the heat value to be cooled is 
practically in proportion to the flow rate of cooling air, it is optimum 
from the standpoint of heat value balance that the centrifugal fan having 
the centrifugal double-impeller structure is constructed to provide the 
ratio W1:W2=2:1 of the flow rate of the first cooling air to the flow rate 
of the second cooling air. By so setting the flow rate ratio, the loss of 
horsepower consumed by the centrifugal fan can be suppressed and a 
reduction in fuel economy can be prevented. Also, since the centrifugal 
fan is not needed to be rotated at a rotational speed in excess of a 
required value, noise can be educed. Further, overcooling of the engine 
can be avoided that may occur due to the excessive flow rate W2 of the 
second cooling air. 
Taking into account variations in an actual engine room caused by a 
difference in resistance against passing air between the flow path on the 
heat exchanger side through which the first cooling air passes and the 
flow path on the engine side through which the second cooling air passes, 
manufacturing errors and so on, however, a proper range of the flow rate 
ratio of the first cooling air to the second cooling air is 
W1:W2=2:0.5-2:1.5. 
In the above engine cooling system, preferably, the centrifugal fan is 
constructed to meet W1:W2=2:1 on an assumption that flow rates of the 
first cooling air and the second cooling air induced by the centrifugal 
fan are W1 and W2, respectively. 
In the above engine cooling system, preferably, the centrifugal fan 
comprises one double-suction impeller made up of a first impeller portion 
fixed to a hub, which is fixed to the rotary shaft, for sucking the first 
cooling air from one end side of the rotary shaft and blowing off the 
first cooling air toward an outer circumference of the fan, and a second 
impeller portion fixed to the hub on the side opposite to the first 
impeller portion for sucking the second cooling air from the other end 
side of the rotary shaft and blowing off the second cooling air toward the 
outer circumference of the fan. 
In the above engine cooling system, preferably, the centrifugal fan 
comprises a first single-suction impeller fixed to a first hub, which is 
fixed to the rotary shaft, for sucking the first cooling air from one end 
side of the rotary shaft and blowing off the first cooling air toward an 
outer circumference of the fan, and a second single-suction impeller fixed 
to a second hub, which is fixed to the rotary shaft, for sucking the 
second cooling air from the other end side of the rotary shaft and blowing 
off the second cooling air toward the outer circumference of the fan. 
Also, preferably, the above engine cooling system further comprises a 
spiral case disposed near an outlet region of the centrifugal fan for 
slowing down the cooling air blown off from the centrifugal fan to restore 
pressure of the blown-off cooling air. 
With this feature, the swirling components contained in streams from 
outlets of the centrifugal fan, which have been all wasted as pressure 
loss in the past, can be restored as a pressure, and therefore the fan 
efficiency can be improved. Hence, the cooling air can be induced at a 
larger flow rate and a higher pressure correspondingly. As a result of 
such a higher pressure of the cooling air, the degree of sealing of the 
engine room can be increased and noise can be further reduced 
correspondingly. 
In the above engine cooling system, preferably, a noise absorbing material 
is attached to at least part of an internal surface of the spiral case. 
This feature is effective in making the engine room quieter as a whole 
because noise produced by the centrifugal fan, which is one of major noise 
sources, can be absorbed by the noise absorbing material. 
Also, preferably, the above engine cooling system further comprises driving 
means for transmitting driving forces to rotate the rotary shaft of the 
centrifugal fan and enabling a rotational speed of the centrifugal fan to 
be set independently of a rotational speed of the engine. 
Specifically, the rotational speed of the centrifugal fan is set 
independently of the rotational speed of the engine by the driving means 
for uniquely generating driving forces from electric energy or hydraulic 
energy, for example, regardless of the engine rotation. This enables the 
centrifugal fan to be rotated at an optimum rotational speed depending on 
working circumstances without being affected by the rotational speed of 
the engine (=rotational speed of a water pump). In work at low 
temperature, for example, if the fan is rotated directly depending on the 
engine rotation, overcooling of the engine may occur due to the excessive 
rotation of the centrifugal fan resulted from securing a usual rated 
rotational speed of the engine. By contrast, in the present invention, 
since the rotational speed of the centrifugal fan can be set to a smaller 
value while maintaining the rotational speed of the engine at the usual 
rated value, the engine can be avoided from lowering its performance due 
to overcooling, and the lower rotational speed of the centrifugal fan can 
prevent an increase in noise generated from the fan. As another example, 
in work performed in highlands, if the fan is rotated directly depending 
on the engine rotation, a reduction in engine performance may occur due to 
deficiency of the cooling ability and resultant overheating of the engine, 
because the engine rotational speed is set to a smaller value to prevent 
stalling of the engine and the rotational speed of the centrifugal fan is 
reduced correspondingly. By contrast, in the present invention, since the 
rotational speed of the engine can be set to a smaller value while 
maintaining the rotational speed of the centrifugal fan as usual, the 
engine can be avoided from lowering its performance due to overheating. 
In the above engine cooling system, preferably, the driving means is any 
one of means for generating the driving forces from electrical energy and 
means for generating the driving forces from hydraulic energy. 
Further, according to the present invention, there is provided a 
construction machine comprising an engine mounted in an engine room, a 
hydraulic pump driven by the engine, actuators driven by a hydraulic fluid 
delivered from the hydraulic pump, and an engine cooling system including 
at least one heat exchanger including a radiator for cooling water to cool 
the engine, and a cooling fan rotated through a rotary shaft for inducing 
cooling air to cool the heat exchanger, the engine, and electric 
equipments disposed around the engine, wherein the cooling fan is a 
centrifugal fan of a centrifugal double-impeller structure for inducing 
first cooling air taken into the engine room through a first air inlet 
port to cool the heat exchanger, and second cooling air taken into the 
engine room through a second air inlet port to cool the engine and the 
electric equipments, and the centrifugal fan is constructed to meet 
W1:W2=2:0.5-2:1.5 on an assumption that flow rates of the first cooling 
air and the second cooling air induced by the centrifugal fan are W1 and 
W2, respectively.

BEST MODE FOR CARRYING OUT THE INVENTION 
Hereunder, embodiments of an engine cooling system of the present invention 
will be described with reference to the drawings. Any of the embodiments 
described below represents an embodiment of the engine cooling system 
installed in an engine room of a hydraulic excavator comprising a 
hydraulic pump driven by an engine and actuators driven by a hydraulic 
fluid delivered from the hydraulic pump. 
First Embodiment 
A first embodiment of the present invention will be described with 
reference to FIGS. 1 to 6. This embodiment embodies the invention as an 
engine cooling system which is installed in an engine room of a hydraulic 
excavator. 
FIG. 1 is a side sectional view showing the structure of an engine room of 
a hydraulic excavator to which this embodiment is applied. 
In FIG. 1, the engine cooling system of this embodiment is installed in an 
engine room 1 in which an engine 5 is mounted. The engine cooling system 
mainly comprises an intercooler 6a for precooling combustion air supplied 
to the engine 5, an oil cooler 6b for cooling a working fluid for the 
hydraulic excavator, a radiator 6c for cooling water to cool the engine 5, 
a centrifugal fan 4 driven by a fan belt 3 to which power is transmitted 
from a crankshaft 2 of the engine 5, and suction tubes 8a, 8b for 
introducing two streams of cooling air 50, 51 (described later) to two 
suction openings of the centrifugal fan 4, respectively. In upper and 
lower walls of the engine room 1, there are formed a first air inlet port, 
e.g., a cooling air inlet port 7a, and a second air inlet port, e.g., 
cooling air inlet ports 7b, for taking ambient air into the engine room, 
as well as an exhaust port 9 for discharging air therefrom. Further, 
electric equipments such as an alternator 10 are installed in the engine 
room 1 near the engine 5. 
FIG. 2 is a view looking from plane II--II as indicated by arrows in FIG. 
1, the view showing the detailed structure of the centrifugal fan 4, FIG. 
3 is a sectional view of presser molds 11a, 11b for shaping the 
centrifugal fan 4, the view corresponding to a section taken along line 
III--III in FIG. 2, and FIG. 4 is a sectional view of the presser molds 
11a, 11b for shaping the centrifugal fan, the view corresponding to a 
section taken along line IV--IV in FIG. 2. In FIGS. 3 and 4, for easier 
understanding of relation between portions of a shaping cavity and 
portions of the centrifugal fan 4, the portions of the shaping cavity are 
denoted by the same reference numerals as the corresponding portions of 
the centrifugal fan 4. 
In FIGS. 1 to 4, the centrifugal fan 4 is of a centrifugal double-impeller 
structure comprising two centrifugal impeller structures. The centrifugal 
fan 4 comprises a hub plate 4a fixed to a rotary shaft 4h, and one 
double-suction impeller 4b fixed to the hub plate 4a. The impeller 4b is 
made up of a first impeller portion 4bL and a second impeller portion 4bR 
each provided in one of both sides of the hub plate 4a and having a 
plurality of vanes. The first impeller portion 4bL includes a rotary 
shroud 4g provided thereon. The centrifugal fan 4 sucks air through 
suction openings 4d, 4e on both sides and blows off the sucked air through 
blowoff openings 4f opened toward an outer circumference of the fan. 
Further, the centrifugal fan 4 is constructed such that a ratio of an 
effective width L1 of the first impeller portion 4bL to an effective width 
L2 of the second impeller portion 4bR is L1:L2=2:0.5-2:1.5. 
The hub plate 4a has a diameter Dh smaller than a suction opening diameter 
D1 of the first impeller portion 4bL provided with the rotary shroud 4g. 
This gives the centrifugal fan 4 a structure capable of being integrally 
molded by using the molds 11a, b. For example, injection molding is 
performed by first placing the hub plate 4a made of an iron core having 
the diameter Dh in the mold 11b through an opening of the mold 11b with 
the diameter D1, mating the mold 11a with the mold 11b, and then injecting 
resin through a not-shown injection port. 
More specifically, the second impeller portion 4bR, parts of the first 
impeller portion 4bL outside the outer diameter Dh of the hub plate 4a, 
and end surfaces 4gR of the rotary shroud 4g facing the engine 5 side can 
be press-formed by the presser mold 11b from the right in FIG. 3, while 
parts of the first impeller portion 4bL inside the hub plate 4a, front 
edges of parts of the first impeller portion 4bL outside the outer 
diameter Dh of the hub plate 4a, and end surfaces 4gL of the rotary shroud 
4g facing the radiator 6c side can be press-formed by the presser mold 11a 
from the left in FIG. 3. Because the centrifugal fan thus has the 
structure capable of being integrally molded, the production cost of the 
fan can be greatly reduced. Also, the provision of the rotary shroud 4g 
reduces the occurrence of turbulence in a flow path on the side of the 
first impeller portion 4bL, improves the fan efficiency, and lessens fan 
noise. 
Returning to FIG. 1, in the engine cooling system constructed as described 
above, there are formed two flow paths of cooling air on the radiator 6c 
side and the engine 5 side looking from the centrifugal fan 4. 
More specifically, the first cooling air 50 passing the flow path on the 
radiator 6c side enters the engine room 1 through the cooling air inlet 
port 7a from the exterior of the engine room 1, and is throttled by the 
suction tube 8a after passing the heat exchangers such as the intercooler 
6a, the oil cooler 6b and the radiator 6c, followed by entering the 
centrifugal fan 4. After that, the first cooling air 50 is blown off 
toward the outer circumference of the centrifugal fan 4 and then 
discharged to the exterior through the exhaust port 9 in the upper wall of 
the engine room 1. 
Also, the second cooling air 51 passing the flow path on the engine 5 side 
enters the engine room 1 through the cooling air inlet ports 7b, 7b from 
the exterior of the engine room 1, and flows around the engine 5, the 
various electric equipments such as the alternator 10, and the oil pan 14 
while cooling them. After that, the second cooling air 51 is throttled by 
the suction tube 8b before entering the centrifugal fan 4, is blown off 
toward the outer circumference of the centrifugal fan 4, and is then 
discharged to the exterior through the exhaust port 9 in the upper wall of 
the engine room 1 together with the first cooling air 50 on the radiator 
6c side. 
In this embodiment constructed as described above, the flow path of the 
second cooling air 51 for cooling the engine 5 and the electric equipments 
such as the alternator 10 can be separated from the flow path of the first 
cooling air 50 for cooling the heat exchangers such as the intercooler 6a, 
the oil cooler 6b and the radiator 6c. Therefore, the cooling air is not 
required to be discharged at a large flow rate from the engine 5 side 
unlike conventional, and the flow rate of the second cooling air 51 on the 
engine 5 side can be set to a smaller value (1/4-3/4 of the flow rate of 
the first cooling air 51 as described later) just required for cooling the 
engine 5 side. Consequently, since the cooling air inlet ports 7b, 7b as 
openings on the engine 5 side can be made smaller than conventional, it is 
possible to increase the degree of sealing on the engine 5 side and to 
reduce noise. 
Also, since the flow path of the second cooling air 51 for cooling the 
engine 5 and the electric equipments such as the alternator 10 has a 
different route from the flow path of the first cooling air 50 for cooling 
the heat exchangers such as the intercooler 6a, the oil cooler 6b and the 
radiator 6c, high-temperature air heated after cooling the heat exchangers 
is avoided from flowing around the electric equipments, etc. unlike 
conventional. Therefore, the effect of cooling the electric equipments, 
etc. can be enhanced and the reliability of the electric equipments, etc. 
can be increased. Further, this results in no need of considering heat 
resistance for the electric equipments, etc. which have to been 
heat-resistant in the past, and hence in reduced cost. Further, the 
enhanced effect of cooling the engine 5 makes it possible to reduce a 
proportion of the cooling ability to be achieved by water cooling in the 
total ability required for cooling the engine, and to reduce the cooling 
ability required for the radiator 6c. Consequently, the size and cost of 
the radiator 6c can be cut down. 
Recently, there has been a tendency for engine rooms to have large 
resistance in cooling flow paths because of such demands as for providing 
the intercooler 6a like the engine room 1 in this embodiment, improving 
the degree of sealing of the engine room for a reduction in noise, and 
making the engine room more compact. In spite of such a tendency, cooling 
air is required to be supplied at a flow rate comparable to conventional. 
This entails the necessity of cooling fans capable of inducing cooling air 
at a larger flow rate and a higher pressure. Bearing the above in mind, 
the engine cooling system of this embodiment employs, as a fan, the 
centrifugal fan 4 which can induce cooling air at a larger flow rate and a 
higher pressure than an axial fan and an angular axial fan and which is 
also advantageous in reducing noise, on an assumption of those fans having 
the same outer diameter and the same rotational speed. This point will now 
be described with reference to FIG. 5. 
FIG. 5 is a graph showing, by way of example, fan characteristics of an 
axial fan and a centrifugal fan on condition that both the fans have the 
same outer diameter and the same rotational speed. In FIG. 5, the 
horizontal axis represents flow rate and the vertical axis represents 
static pressure. Characteristic curves denoted by "axial fan" and 
"centrifugal fan" indicate characteristics of the axial fan and the 
centrifugal fan set alone (i.e., a characteristic of each sole fan 
measured with the fan not disposed in a flow path). Two resistance curves 
(1) and (2) represent characteristics of cooling flow paths alone in 
engine rooms (i.e., characteristics determined uniquely by the structures 
of the flow paths). Then, intersect points between the fan characteristic 
curves and the resistance curves indicate working points resulted when the 
relevant fan is disposed in the relevant flow path, and provide the 
pressure and flow rate resulted in respective cases. Note that, of the 
resistance curves (1) and (2), the resistance curve (1) represents 
resistance of the cooling flow path in a conventional engine room, and the 
resistance curve (2) represents a characteristic of the cooling flow path 
in a recent engine room designed to meet the demands for providing an 
intercooler, improving the degree of sealing of the engine room for a 
reduction in noise, and making the engine room more compact. 
First, in the case of using the axial fan in the conventional engine room, 
the resultant flow rate and static pressure are Qprop1 and Pprop1, 
respectively, as indicated by the intersect point A between the 
characteristic curve denoted by "axial fan" and the resistance curve (1). 
On the contrary, in the case of using the centrifugal fan in the 
conventional engine room, the resultant flow rate and static pressure, are 
Qturbo1 and Pturbo1, respectively, as indicated by the intersect point B 
between the characteristic curve denoted by "centrifugal fan" and the 
resistance curve (1). Thus, with the same outer diameter and the same 
rotational speed, the centrifugal fan has properties of enabling cooling 
air to be induced at a higher pressure and a larger flow rate than the 
axial fan due to the centrifugal action (details will be described later). 
If the conventional axial fan is applied as it is to the recent engine 
room, the resultant flow rate and static pressure are Qprop2 and Pprop2, 
respectively, as indicated by the intersect point C between the 
characteristic curve denoted by "axial fan" and the resistance curve (2). 
While the static pressure is raised from Pprop1 resulted above when the 
conventional axial fan is applied to the conventional engine room, thereby 
enabling cooling air to be induced at a higher pressure, the flow rate is 
reduced from Qprop1 resulted above when the conventional axial fan is 
applied to the conventional engine room. In order to provide a flow rate 
comparable to Qprop1 resulted in the conventional engine room, therefore, 
the rotational speed must be increased, which necessarily magnifies noise 
to a large extent. On the other hand, when the centrifugal fan is applied 
to the recent engine room, the resultant flow rate and static pressure can 
be given as Qturbo2 (=Qprop1) and Pturbo2, respectively, as indicated by 
the intersect point D between the characteristic curve denoted by 
"centrifugal fan" and the resistance curve (2). It is thus possible to 
provide almost the same flow rate as Qprop1, i.e., the flow rate resulted 
when the axial fan is applied to the conventional engine room, and to 
raise the static pressure twice or more Pprop1, i.e., the static pressure 
resulted when the axial fan is applied to the conventional engine room. 
Those characteristics of the centrifugal fan can be explained, by way of 
example, as follows. 
Generally, a theoretical pressure rise Pth of a fan is expressed by the 
following formula: 
EQU Pth=P(u.sub.2.sup.2 -u.sub.1.sup.2)/2+P(v.sub.2.sup.2 
-v.sub.1.sup.2)/2+P(w.sub.2.sup.2 -w.sub.1.sup.2)/2 
where u is the circumferential speed of the fan, v is the absolute speed of 
the flow, w is the relative speed of the flow, and suffixes 1, 2 represent 
that corresponding values are measured at an inlet and an outlet of the 
fan, respectively. 
In the above formula, the first term of the right side, i.e., 
P(u.sub.2.sup.2 -u.sub.1.sup.2)/2, represents a pressure rise due to the 
effect of centrifugal forces, the second term of the right side, i.e., 
P(v.sub.2.sup.2 -v.sub.1.sup.2)/2, represents change in kinetic energy 
(rise of dynamic pressure), and the third term of the right side, i.e., 
P(w.sub.2.sup.2 -w.sub.1.sup.2)/2, represents a pressure rise due to the 
effect of slowdown in the flow path. Looking now at the first term, the 
value of the first term for an axial fan is equal to zero (0) because the 
axial fan has an inlet and an outlet of the same diameter and hence 
u.sub.1 -u.sub.2. For a centrifugal fan, however, the effect of 
centrifugal forces due to the first term is maximally developed because 
the centrifugal fan has a fan outlet larger than a fan inlet. Accordingly, 
the centrifugal fan can induce cooling air at a higher pressure than the 
axial fan, and hence can easily supply the cooling air at a larger flow 
rate. Note that while the characteristic of a centrifugal fan has been 
explained above in comparison with that of an axial fan, the above 
explanation is also equally applied to comparison with an angular axial 
fan. 
As described above, by using the centrifugal fan 4 as a cooling fan, 
cooling air can be induced at a larger flow rate and a higher pressure 
than the axial fan and the angular axial fan with the same outer diameter 
and the same rotational speed. Accordingly, even with intent to induce 
cooling air at a larger flow rate and a higher pressure to provide a flow 
rate in the recent engine room comparable to conventional, the rotational 
speed is not required to be increased unlike the axial fan and the angular 
axial fan, and noise can be reduced. Further, since the rotary shroud 4g 
is provided on the first impeller portion 4bL, the cooling air is 
prevented from leaking radially through gaps between the suction tube 8a 
and the first impeller portion 4bL, thus enabling the fan efficiency to be 
improved. Noise can be further reduced correspondingly. 
Also, in the conventional structure using an axial fan, because cooling air 
blown off from the axial fan cools an engine disposed downstream of the 
fan, rotation of the fan causes a stream of cooling air, including a 
swirling component, to strike against complicated shapes defined by the 
engine and various members surrounding it, thereby locally producing 
reversed streams of the cooling air. These reversed streams impede the 
cooling air from flowing smoothly. By contrast, in the cooling system of 
this embodiment, since the engine 5 is disposed upstream of the 
centrifugal fan 4, the second cooling air 51 including no swirling 
components flows toward the centrifugal fan 4 along the surface of the 
engine 5 and the surface of the oil pan 14 below the engine 5. As a 
result, the occurrence of reversed streams produced in the conventional 
system can be held down. 
Further, by setting the ratio of the effective width L1 of the first 
impeller portion 4bL to the effective width L2 of the second impeller 
portion 4bR to L1:L2=2:0.5-2:1.5, a ratio of the flow rate of the first 
cooling air 50 to the flow rate of the second cooling air 51 is given by 
W1:W2=2:0.5-2:1.5. This is effective in optimizing flow rate balance 
between the first cooling air 50 and the second cooling air 51. That point 
will be described with reference to FIG. 6. 
FIG. 6 shows percentages of respective heat values produced as horsepower 
and consumed as various losses on an assumption that the total heat value 
generated by an engine is 100%. As shown, it is generally known that, of 
the total heat value 100% generated by an engine, the percentage of the 
heat value consumed by cooling of radiator water, for example, is about 
30%, and the percentage of the heat value consumed as exhaust heat and 
radiation heat from outer walls of the engine is about 33% (Internal 
Combustion Engine Handbook, by Asakura Shoten, Publisher). Then, what 
percentage of the latter 33% is occupied by radiation heat can be 
calculated, by way of example, as follows. 
In general, the relationship between the total heat value and horsepower of 
an engine is expressed by the following basic formula: 
EQU Q=Ne.times.b.times.Hu/60 
where 
Q: heat value discharged by cooling water kcal/min! 
Ne: horsepower PS! 
b: fuel consumption rate kg/PSHr! 
Hu: low heat value generated by fuel (=10500 kcal/kg!) 
Taking a typical fuel consumption rate of P.ltoreq.170 g/PSh!, for 
example, the total heat value Q generated by the engine is calculated 
below from the above basic formula: 
EQU P:Q.sub.p =10500 kcal/kg!170.times.10.sup.-3 kg/PSh!.times.135 
PS!=240975 kcal/h! 
Then, because the percentage of the exhaust heat and the radiation heat is 
33% as mentioned above, the heat value consumed as the exhaust heat and 
the radiation heat is given by: 
EQU (exhaust heat+radiation heat)=240975.times.0.33=79522 kcal/h!(1) 
On the other hand, the exhaust heat discharged through a silencer is 
expressed below on condition that the exhaust temperature is 300.degree. 
C., the ambient air temperature is 20.degree. C., the flow rate is 313 
m.sup.3 /s!, and the specific weight is 0.596 kg/m.sup.3 !: 
##EQU1## 
Accordingly, from (1) and (2), the radiation heat is given by: 
##EQU2## 
This value corresponds to about 14.2% of above Q (=240975 kcal/h!). Thus, 
of the total heat value 100% generated by the engine, the percentage of 
the heat value consumed by cooling of radiator water is about 30% and the 
percentage of the heat value consumed as exhaust heat from outer walls of 
the engine is about 14.2%. 
Applying the above consideration to the engine 5 of this embodiment, the 
former percentage practically corresponds to the heat value cooled by the 
first cooling air 50 in the radiator 6c, and the latter percentage 
practically corresponds to the heat value cooled by the second cooling air 
51 around the engine. Therefore, on an assumption that the heat value to 
be cooled is practically in proportion to the flow rate of cooling air, it 
is optimum from the standpoint of heat value balance that the centrifugal 
fan 4 having the centrifugal double-impeller structure is constructed to 
provide the ratio W1:W2=2:1 of the flow rate of the first cooling air 50 
to the flow rate of the second cooling air 51. By so setting the flow rate 
ratio, the loss of horsepower consumed by the centrifugal fan can be 
suppressed and a reduction in fuel economy can be prevented. Also, since 
the centrifugal fan is not needed to be rotated at a rotational speed in 
excess of a required value, noise can be educed. Further, overcooling of 
the engine can be avoided that may occur due to the excessively large flow 
rate W2 of the second cooling air 51. 
Taking into account variations in an actual engine room caused by a 
difference in resistance against passing air between the flow path on the 
heat exchangers 6a-c side through which the first cooling air 50 passes 
and the flow path on the engine 5 side through which the second cooling 
air 51 passes, manufacturing errors and so on, however, a proper range the 
flow rate ratio of the first cooling air 50 to the second cooling air 51 
has of W1:W2=2:0.5-2:1.5. 
Accordingly, in this embodiment, by setting the ratio of the effective 
width L1 of the first impeller portion 4bL to the effective width L2 of 
the second impeller portion 4bR to L1:L2=2:0.5-2:1.5, flow rate balance 
between the first cooling air 50 and the second cooling air 51 can be 
optimized. 
While the first embodiment is designed such that one of the two cooling air 
flow paths (i.e., the first cooling air 50) is allocated to cool the heat 
exchangers such as the intercooler 6a, the oil cooler 6b and the radiator 
6c, and the other cooling air flow path (i.e., the second cooling air 51) 
is allocated to cool the electric equipments such as the alternator 10, 
the engine 5 and the oil pan 14, those objects to be cooled may be 
distributed to the two cooling air flow paths in any other suitable way 
than above. For example, the intercooler 6a of the heat exchangers may be 
disposed downstream of the suction opening 7b on the engine 5 side. In 
essence, the similar advantages as in the above first embodiment can be 
obtained if at least one heat exchanger is disposed in the cooling air 
flow path not on the engine 5 side. 
Also, while the rotary shroud 4g is provided on the first impeller portion 
4bL in the first embodiment, the rotary shroud 4g is not always required 
to be provided just for the purpose of improving the effect of cooling the 
electric equipments, etc. As an alternative, rather than providing the 
rotary shroud 4g on only the first impeller portion 4bL, a separate rotary 
shroud may be additionally provided on the second impeller portion 4bR as 
well so that both the first impeller portion 4bL and the second impeller 
portion 4bR have rotary shrouds. This modification can improve the fan 
efficiency and reduce noise. 
Further, while in the above first embodiment the centrifugal fan 4 has the 
double-suction impeller 4b which sucks air from two directions and blows 
off the air in one direction, the centrifugal fan is not limited to such a 
structure, but may comprise two single-suction impellers arranged in 
back-to-back relation. This modified embodiment is shown in FIG. 7. 
Equivalent members to those in the first embodiment are denoted by the 
same reference numerals. 
FIG. 7 is a side sectional view showing the structure of an engine room in 
which is installed an engine cooling system according to the modified 
embodiment including two single-suction impellers. This modified 
embodiment differs from the first embodiment shown in FIG. 1 in that a 
centrifugal fan 104 made up of two single-suction impellers 104Lb, 104Rb 
with their suction openings facing away from each other in opposite 
directions is provided in place of the centrifugal fan 4 made up of one 
double-suction impeller 4b. The centrifugal fan 104 comprises a first hub, 
e.g., a hub plate 104La, fixed to the rotary shaft 4h, a first impeller 
104Lb fixed to the hub plate 104La, a second hub, e.g., a hub plate 104Ra, 
fixed to the rotary shaft 4h, a rotary shroud 104Lg fixed to the first 
impeller 104Lb on the suction side, a second impeller 104Rb fixed to the 
hub plate 104Ra, and a rotary shroud 104Rg fixed to the second impeller 
104Rb on the suction side. Then, the first impeller 104Lb on the left side 
as viewed in the drawing sucks the first cooling air 50 taken in through 
the cooling air inlet port 7a and having cooled the intercooler 6a, the 
oil cooler 6b, the radiator 6c, etc., and blows off the sucked air toward 
the exhaust port 9. The second impeller 104Rb on the right side as viewed 
in the drawing sucks the second cooling air 51 taken in through the 
cooling air inlet ports 7b and having cooled the engine 5, the electric 
equipments such as the alternator 10, and the oil pan 14, and blows off 
the sucked air toward the exhaust port 9. 
The other construction is substantially the same as in the first 
embodiment. 
This modified embodiment can also provide the similar advantages as in the 
first embodiment. 
In addition to those advantages, when the present invention is applied to 
the conventional structure using a centrifugal fan made up of one 
single-suction impeller, the impeller can be reused in fabricating the 
centrifugal fan according to the above modified embodiment. For this 
reason, the above modified embodiment can be more easily practiced than 
the first embodiment in which the impeller 4b made up of the first and 
second impeller portions 4bL, 4bR must be newly manufactured. Another 
advantage is that since the first and second single-suction impellers 
104Lb, Rb can be each uniquely selected in its outer diameter, the number 
of vanes, etc. without limitations imposed from mutual relation, the 
degree of freedom in design can be increased. 
Second Embodiment 
A second embodiment of the present invention will be described with 
reference to FIGS. 8 and 9. In this embodiment, a spiral case is provided 
in addition to the construction of the first embodiment. Equivalent 
members to those in the first embodiment are denoted by the same reference 
numerals. 
FIG. 8 is a side sectional view showing the structure of the engine room 1 
in which an engine cooling system according to this embodiment is 
installed. This embodiment differs from the first embodiment in that a 
spiral case 212 is provided to cover an outlet region of the centrifugal 
fan 4 in the vicinity thereof. The spiral case 212 has a structure 
dividable into a plurality of parts along a plane containing the rotary 
shaft of the centrifugal fan 4, and is attached in place after mounting 
the engine 5, the intercooler 6a, the oil cooler 6b, the radiator 6c, the 
centrifugal fan 4, etc. 
FIG. 9 is a perspective view showing the detailed structure of the spiral 
case 212 and thereabout. The spiral case 212 defines a flow path therein 
which has a shape gradually increasing a cross-sectional area of the flow 
path. 
In the construction shown in FIGS. 8 and 9, as with the first embodiment, 
the first cooling air 50 having cooled the intercooler 6a, the oil cooler 
6b, the radiator 6c, etc. and the second cooling air 51 having cooled the 
engine 5, the oil pan 14 and the electric equipments such as the 
alternator 10 are sucked into the centrifugal fan 4 and blown off 
therefrom. After that, the air is gradually slowed down in the spiral case 
212 while restoring pressure, and then discharged through the exhaust port 
9. 
The other structure is substantially the same as in the first embodiment. 
This embodiment can also provide the similar advantages as in the first 
embodiment. In addition to those advantages, since the swirling components 
contained in streams from outlets of the centrifugal fan 4, which have 
been all wasted as pressure loss in the past, can be restored as a 
pressure with the provision of the spiral case 212, it is possible to 
improve the fan efficiency and to induce the cooling air at a larger flow 
rate and a higher pressure correspondingly. As a result of such a higher 
pressure of the cooling air, the degree of sealing of the engine room 1 
can be improved and noise can be further reduced correspondingly. 
Third Embodiment 
A third embodiment of the present invention will be described with 
reference to FIG. 10. In this embodiment, a noise absorbing material is 
added to the construction of the second embodiment. Equivalent members to 
those in the second embodiment are denoted by the same reference numerals. 
FIG. 10 is a side sectional view showing the structure of the engine room 1 
in which an engine cooling system according to this embodiment is 
installed. This embodiment differs from the second embodiment in that a 
noise absorbing material 313 is attached to a part of the inner surface of 
the spiral case 212. As an alternative, the noise absorbing material 313 
may be attached to the entire inner surface of the spiral case rather than 
a part thereof. 
The other structure is substantially the same as in the second embodiment. 
In addition to the similar advantages as in the second embodiment, this 
embodiment is also effective in making the engine room 1 quieter as a 
whole because noise produced by the centrifugal fan 4, which is one of 
major noise sources, can be absorbed by the noise absorbing material 313. 
Fourth Embodiment 
A fourth embodiment of the present invention will be described with 
reference to FIG. 11. In this embodiment, a centrifugal fan is driven by a 
driving source different from the engine crankshaft. 
FIG. 11 is a side sectional view showing the structure of an engine room in 
which an engine cooling system according to this embodiment is installed. 
This embodiment differs from the first embodiment particularly in that a 
centrifugal fan 403 is driven by an electric motor 406 (described later). 
It is to be noted that the other structure is also somewhat different in 
various components from the first to third embodiments, and therefore the 
following description will be made on all members, including equivalent 
members to those in the first to third embodiments. The engine cooling 
system is installed in an engine room 401 in which an engine 402 is 
mounted. The engine cooling system mainly comprises an intercooler 409a 
for precooling combustion air supplied to the engine 402, an oil cooler 
409b for cooling a working fluid for a hydraulic excavator, a radiator 
409c for cooling water to cool the engine 402, a centrifugal fan 403 
having a rotary shaft 403a rotatably supported by a bearing 413 which is 
in turn supported by the engine 402 through a bearing support member 423, 
driving means, e.g., an electric motor 406 to generate driving power from 
electric energy, for transmitting driving forces to the rotary shaft 403a 
of the centrifugal fan 403 through a fan belt 422, suction tubes 410a, 
410b for introducing two streams of cooling air 450, 451 (described later) 
to two suction openings 403d, 403e of the centrifugal fan 403 located on 
both sides (on the left and right sides as viewed in the drawing), 
respectively, and a water pump 405 for circulating the cooling water of 
the engine 402 into the radiator 409c through cooling water tubes 415. 
The water pump 405 is driven by the power transmitted from a crankshaft 411 
of the engine 402 to a rotary shaft (water pump rotary shaft) 421 through 
a belt 412. 
The centrifugal fan 403 is of a centrifugal double-impeller structure 
comprising two centrifugal impeller structures. The centrifugal fan 403 
comprises a hub plate (not shown) fixed to a rotary shaft 403a, and one 
double-suction impeller 403b fixed to the hub plate. The impeller 403b is 
made up of a first impeller portion 403bL and a second impeller portion 
403bR each provided in one of both sides of the hub plate and having a 
plurality of vanes. The centrifugal fan 403 sucks air through the suction 
openings 403d, 403e on both sides and blows off the sucked air through 
blowoff openings 403f opened toward an outer circumference of the fan. 
In upper and lower walls of the engine room 401, there are formed cooling 
air inlet ports 407a, 407b for taking in ambient air, and an exhaust port 
408 for discharging air. Though not specifically shown, electric 
equipments such as an alternator are installed in the engine room 401 near 
the engine 402. 
In the engine cooling system constructed as described above, there are 
defined two flow paths of cooling air on the radiator 409c side and the 
engine 402 side looking from the centrifugal fan 403. 
More specifically, the first cooling air 450 passing the flow path on the 
radiator 409c side enters the engine room 401 through the cooling air 
inlet port 407a from the exterior of the engine room 401, and is throttled 
by the suction tube 410a after passing the heat exchangers such as the 
intercooler 409a, the oil cooler 409b and the radiator 409c, followed by 
entering the centrifugal fan 403. After that, the first cooling air 450 is 
blown off toward the outer circumference of the centrifugal fan 403 and 
then discharged to the exterior through the exhaust port 408 in the upper 
wall of the engine room 401. 
Also, the second cooling air 451 passing the flow path on the engine 402 
side enters the engine room 401 through the cooling air inlet ports 407b 
from the exterior of the engine room 401, and flows around the engine 402 
and the oil pan 404, etc. disposed near and below the engine 402, while 
cooling them. After that, the second cooling air 451 is throttled by the 
suction tube 410b before entering the centrifugal fan 403, is blown off 
toward the outer circumference of the centrifugal fan 403,; and is then 
discharged to the exterior through the exhaust port 408 in the upper wall 
of the engine room 401 together with the first cooling air 450 on the 
radiator 409c side. 
Moreover, the first impeller portion 403bL and the second impeller portion 
403bR are constructed such that a flow rate ratio of the first cooling air 
450 to the second cooling air 451 is W1:W2=2:0.5-2:1.5. 
In addition to the similar advantages as in the first embodiment, this 
embodiment constructed as described above can also provide advantages 
below. 
First, since the rotational speed of the centrifugal fan 403 can be 
uniquely set by the electric motor 406 independently of the rotational 
speed of the engine 402, the centrifugal fan 403 can be rotated at an 
optimum rotational speed depending on working circumstances without being 
affected by the rotational speed of the engine 402 (=rotational speed of 
the water pump 405). In work at low temperature, for example, the 
conventional structure where the fan is rotated directly depending on the 
engine rotation may cause overcooling of the engine due to the excessive 
rotation of the centrifugal fan resulted from securing a usual rated 
rotational speed of the engine. By contrast, in this embodiment, since the 
rotational speed of the centrifugal fan 403 can be set to a smaller value 
while maintaining the rotational speed of the engine 402 at the usual 
rated value, the engine 402 can be avoided from lowering its performance 
due to overcooling, and the lower rotational speed of the centrifugal fan 
403 can prevent an increase in noise generated from the fan. As another 
example, in work performed in highlands, the conventional structure where 
the fan is rotated directly depending on the engine rotation may cause a 
reduction in engine performance due to deficiency of the cooling ability 
and resultant overheating of the engine, because the engine rotational 
speed is set to a smaller value to prevent stalling of the engine and the 
rotational speed of the centrifugal fan is reduced correspondingly. By 
contrast, in this embodiment, since the rotational speed of the engine 402 
can be set to a smaller value while maintaining the rotational speed of 
the centrifugal fan 403 as usual, the engine 402 can be avoided from 
lowering its performance due to overheating. 
Also, since the rotational speed of the engine 402 is set regardless of the 
rotational speed of the centrifugal fan 403, an advantage of improving 
fuel economy can also be obtained. 
The weight of the centrifugal fan 403 may be possibly larger than that of 
the conventional axial fan. In this embodiment, however, since the 
centrifugal fan 403 is not fixed to the rotary shaft 421 of the engine 
402, but driven by the electric motor 406, the load imposed on a bearing 
(though not shown, installed in the engine 402) of the rotary shaft 421 of 
the engine 402 is not increased. In other words, the larger weight of the 
centrifugal fan 403 can be borne by designing the bearing 413 of the fan 
rotary shaft 403a so as to withstand the larger weight case by case. 
Therefore, the engine cooling system of this embodiment can also be easily 
applied even to an existing engine which has been designed in anticipation 
of the use of an axial fan, just by newly providing the bearing support 
member 423 and the bearing 413. 
Further, the fan center position (=position of the fan rotary shaft) 
optimum for cooling the intercooler 409a, the oil cooler 409b and the 
radiator 409c depends on the sizes and shapes of those three heat 
exchangers. In the conventional structure where the fan rotary shaft is 
directly rotated by the engine crankshaft, a difficulty in shifting the 
fan rotary shaft with respect to the engine position may inevitably 
entails an offset of the fan rotary shaft from an optimum position and 
give rise to a reduction in the flow rate of cooling air. By contrast, in 
this embodiment, since the fan rotary shaft 403a can be set in any desired 
position, it is possible to shift the fan rotary shaft 403a to the optimum 
position depending on the three heat exchangers 409a-c and to prevent a 
reduction in the flow rate of cooling air. 
While the centrifugal fan 403 made up of one double-suction impeller 403b 
is used in the above fourth embodiment, a centrifugal fan made up of two 
single-suction impellers with their suction openings facing away from each 
other in opposite directions may be used in place of the centrifugal fan 
403 as with the modified embodiment shown in FIG. 7. This case can also 
provide the similar advantages. 
Also, while the above fourth embodiment has been illustrated as driving the 
impeller 403b by the electric motor 406, a hydraulic motor driven with 
hydraulic energy (hydraulic fluid) may be used in place of the electric 
motor 406. This case can also provide the similar advantages. 
Further, while the above fourth embodiment uses the electric motor 406 
which is a driving source totally independent of the rotation of the 
engine 402, it is not limited to the use of such an independent driving 
source, but may be modified as follows. Specifically, the centrifugal fan 
403 may be rotated at a proper rotational speed depending on working 
circumstances by providing rotational speed changing means which receives 
the rotation of the engine 402 transmitted thereto, changes the input 
rotational speed at a desired speedup/slowdown ratio, and then outputs the 
resultant rotational speed to the centrifugal fan 403. The rotational 
speed changing means may be made up of, e.g., a speed reducing gear 
mechanism having plural types of gears with the different numbers of teeth 
from each other, and speedup/slowdown ratio control means for controlling 
the speedup/slowdown ratio of transmitted rotation in the speed reducing 
gear mechanism, e.g., a gear shifting mechanism for shifting gear 
positions in the speed reducing gear mechanism. These cases can also 
provide the similar advantages. 
In the above first to fourth embodiments, the intercoolers 6a, 409a, the 
oil coolers 6b, 409b and the radiators 6c, 409c are provided as examples 
of heat exchangers. However, the heat exchangers are not limited to those 
examples, but may include any other one(s) such as a condenser for an air 
conditioner. These cases can also provide the similar advantages. 
Moreover, the above first to fourth embodiments have been described in 
connection with, by way of example, the engine cooling system equipped on 
construction machines. However, the present invention is not limited to 
such an application, but also applicable to engine cooling systems 
equipped on automobiles, agricultural machines and other types of 
machines. These cases can also provide the similar advantages. 
INDUSTRIAL APPLICABILITY 
According to the present invention, since the cooling flow path of the 
second cooling air for cooling the engine side can be separated from the 
cooling flow path of the first cooling air for cooling the heat 
exchangers, the size of an opening to be formed on the engine side can be 
set smaller than conventional. Accordingly, the degree of sealing on the 
engine side can be improved and noise can be reduced. Also, since 
high-temperature air heated after cooling the heat exchangers is prevented 
from flowing around the electric equipments, etc. unlike conventional, it 
is possible to enhance the effect of cooling the electric equipments, etc. 
and to improve the reliability of the electric equipments, etc.