Vortex flow blower

A vortex flow blower including a blower casing having an annular flow passageway extending from an inlet port for receiving fluid to an outlet port for discharging the fluid, the outlet port being disposed adjacent to the inlet port, and an impeller accommodated in the blower casing for producing a vortex flow of the fluid in the annular flow passageway. The vortex flow blower is configured for enabling at least one of noise reduction pressure increase and reduction of power requirements of the vortex flow blower, by providing at least one of sectional area reducer for reducing a sectional area of the annular flow passageway which annular flow passageway includes an annular groove disposed in facing relation to vanes of the impeller, and a partition wall partitioning a part of the circumference of the annular groove so that the inlet port and the outlet port being provided at opposite end portions of the annular groove partitioned by the partition wall with the sectional area reducer is disposed at a position of the annular passageway located between the outlet port of the annular passageway and a midpoint between the inlet port and the outlet port of the annular passageway, and an auxiliary flow supply path for supplying an auxiliary flow of the fluid introduced to the annular flow passageway from the inlet port so as to conduct the fluid in a direction to form the vortex flow.

BACKGROUND OF THE INVENTION 
The present invention relates to the construction of a vortex flow blower 
for improving performance thereof when such vortex flow blower is operated 
as a centrifugal gas pump or a centrifugal fluid pump such as a WESCO 
pump. 
This type of vortex blower as a centrifugal pump is generally provided with 
an impeller having a large number of vanes and disposed in an annular flow 
path, an inlet or suction port and an outlet or discharge port both 
communicate with the interior of the annular flow path, and a partition 
wall for partitioning the section from the discharge port to the suction 
port through a very small gap with respect to a vane passing path. Gas or 
liquid (both will be generically called "fluid" hereinafter) which has 
been introduced from the suction port is rotated and pressurized in the 
form of a vortex flow in the annular flow path by rotating the impeller 
disposed in the same path, and the fluid is then discharged from the 
discharge port. 
In prior art single-side impeller type centrifugal pumps, as described for 
example in Laid-Open Japanese Patent Application No. 51-70512, an annular 
groove in a casing is made nearly in a semi-elliptical form expressed by 
d&lt;(D2-D1)/4 where D2 is the outside diameter of the annular groove, D1 is 
inside diameter, and d is depth, thereby preventing the reverse flow of a 
fluid. It is, therefore, possible to provide a small-volume but 
high-static pressure centrifugal blower. 
Also, in prior-art double-side blade-type centrifugal pumps, the annular 
groove of the casing is provided with an annular projection on the outer 
periphery of the annular groove as described in Laid-Open Japanese Patent 
Application No. 49-135209, thereby improving pump output performance by 
preventing occurrence of a breakaway flow. 
The prior art described above are concerned with producing a high static 
pressure within a range of small air volume, and a construction required 
for noise reduction is not taken into consideration. Furthermore, such 
above-described prior art having projections and shallow grooves provided 
all around the annular groove, have such a problem as a decrease in the 
sectional area of a portion extending from a suction port of the annular 
groove to a suction-discharge center and accordingly a decrease in the air 
volume. This decrease in the sectional area of this position hardly 
contributes toward the noise reduction. 
Additionally, while the above-described type of a centrifugal pump is 
relatively easy to handle and therefore is utilized in various fields 
there is another problem related to such structure. More particularly, the 
impeller rotates continuously and the fluid which has been introduced from 
the suction port is rotated and pressurized in the form of a vortex flow 
in the annular flow path, then is carried to the discharge port by the 
action of the partition wall. At this time, as, for example, disclosed in 
Japanese Utility Model Laid Open No. 91308/76, a portion of the fluid is 
allowed to remain between adjacent vanes of the impeller and is thereby 
conveyed to the suction portion side, which fluid portion will hereinafter 
be referred to as "carry-over flow". The carry-over flow passes the 
partition wall and is conveyed to the suction side while the rotation 
thereof is suppressed. On the suction side, the pressurized carry-over 
flow is released throughout the entire vane width and expands 
substantially uniformly in the flow path. As a result, the amount of fluid 
which is introduced decreases accordingly, that is, an effective amount of 
fluid conveyed decreases, and hence the characteristic thereof remains 
poor. 
As mentioned above, the fluid of a large rotation and high pressure on the 
discharge port side flows out from the discharge port. But according to an 
analysis made by the present inventors, it turned out that the carry-over 
flow not only decreases an effective amount of fluid conveyed, but also 
operates disadvantageously in the following point. Once the carry-over 
flow of high pressure is released on the suction port side, it is released 
throughout the entire vane width in this position and expands 
substantially uniformly without rotation in the flow path. As a result, 
this expanded flow is mixed with fluid introduced from the suction port 
without changing the length of wetted perimeter and causes disturbance in 
the fluid introduced from the suction port. Due to this disturbance, the 
fluid introduced from the exterior through the suction port cannot form a 
rotating flow in the flow path portion near the suction port, and only 
after passing this mixing region, it forms an effective rotating flow. 
According to an experimental measurement made by the present inventors, 
this mixing region was about 40.degree. in terms of the angle of 
circumference from the suction port to the discharge port side. In the 
conventional centrifugal pumps, therefore, a rotating flow cannot be 
formed at an angle corresponding to such mixing region, i.e., about 
40.degree., so it is impossible to raise the pressure and hence the 
pressure is low. It became clear that this had a bad influence on the 
improvement of characteristics and also became clear that such disturbance 
badly affected the generation of noise. 
It is a well-known fact that the disturbance of fluid causes the 
deterioration of performance also in hydraulics and aerodynamics. 
SUMMARY OF THE INVENTION 
It is an object of the present invention to provide a vortex flow blower 
having improved performance. 
It is another object of the present invention to provide a vortex flow 
blower which is able to lower noise level and to obtain a high static 
pressure over the entire range of air volume. 
It is a further object of the present invention to provide a vortex flow 
blower capable of forming a rotating flow more effectively throughout the 
entirety of a flow path. 
It is still a further object of the present invention to provide a vortex 
flow blower capable of forming a rotating flow smoothly from the vicinity 
of a suction port. 
It is yet another object of the present invention to provide a vortex flow 
blower capable of utilizing a carry-over flow more effectively. 
It is a further object of the present invention to provide a vortex flow 
blower capable of diminishing an influent loss of fluid. 
According to the present invention, a vortex flow blower such as a 
centrifugal pump includes an impeller and a casing which is provided with 
an inlet or suction port and an outlet or discharge port and houses the 
impeller, and has an annular groove provided between the suction port and 
the discharge port along the direction of rotation of the impeller, in a 
part facing to blades of the impeller in the casing, and the annular 
groove has the sectional area thereof reduced in a part of a zone 
extending between the discharge or outlet port and point midway between 
the inlet or suction port and the discharge or outlet port of the annular 
groove, thereby enabling noise reduction and high static pressure. 
According to a feature of the present invention, a section of the annular 
groove to be reduced is cut on a plane passing a rotating shaft and formed 
of a slanting projection extending from the vicinity of the outer 
peripheral edge of the annular groove to the bottom of the annular groove. 
In accordance with another feature of the present invention, the depth of 
the annular groove from the surface of casing facing the impeller 
increases in the order of an intermediate position between the central 
part of the annual passage and the suction port, the central part of the 
annual passage, and an intermediate position between the central part of 
the annular passage and the discharge port. 
In accordance with the present invention, the impeller is driven to rotate 
by a primer mover, producing an internal flow of a fluid flowing out from 
the outer peripheral section, and the reduced area of the annular groove 
provides a slant face to the internal flow of the fluid flowing out from 
the impeller, guiding the fluid to the inner periphery so as to positively 
change the course of the internal flow. Therefore, the internal flow of 
fluid flowing out from the outer peripheral section of the impeller is 
guided to the inner peripheral section, flowing close to the flow of fluid 
flowing out from a portion spaced from the outer peripheral section of the 
impeller. In this manner, the occurrence of breakaway of the fluid which 
is likely to be caused by a difference in flow velocity between the fluid 
flowing out from the outer peripheral section of the impeller and the 
fluid flowing from the position spaced from the outer peripheral section 
is minimized, thereby enabling controlling occurrence of sound and, at the 
same time, controlling a loss resulting from internal flow turbulence. 
Thus it is possible to obtain a high static pressure. Furthermore, it is 
possible to control the occurrence of noise by guiding, in the vicinity of 
a no-discharge operation, the internal flow rapidly into the inner 
peripheral section and by decreasing the inflow velocity of the fluid at 
the inner peripheral section of the impeller and, at the same time, it is 
possible to control a loss resulting from the internal flow turbulence, 
thereby obtaining an increased static pressure. Additionally, it is 
possible to prevent a decrease in the air volume because the area of flow 
passage is kept unchanged on the suction or inlet side. 
According to another feature of the present invention, the vortex flow 
blower such as a centrifugal pump is provided with an auxiliary flow 
supply path for supplying an auxiliary flow to fluid introduced from a 
suction or inlet port to conduct the fluid in a direction to form a 
rotating flow in a flow path. 
In accordance with the present invention, the auxiliary flow may be fed 
from the exterior, but it is desirable and advantageous to utilize a 
carry-over flow. Furthermore, it is desirable that the auxiliary flow be 
supplied forwards relative to an advancing direction of an impeller, more 
specifically, at an angle in the range from 5.degree. to 35.degree., using 
as a reference plane the surface of the partition wall of the flow passage 
which defines a very small gap with respect to the impeller. 
In connection with utilizing a carry-over flow as the auxiliary flow, the 
present invention utilizes a discharge guide portion for guiding the 
carry-over flow so as to be discharged obliquely forwards relative to the 
advancing direction of the vanes of the impeller, on the suction or inlet 
port side of the partition wall. The partition wall is provided with a 
flow guide portion for conducting the flow from the suction port 
efficiently into the annular flow path. Fluid remaining between adjacent 
vanes is carried to the suction port side in a closed state of a discharge 
or outlet port by the flow guide portion. Although the discharge guide 
portion may be provided separately from the partition wall, it is 
desirable to form it in the flow guide portion of the partition wall. The 
portion of the partition wall where the flow guide portion is to be formed 
may be cut-out in the form of a hole or may be cut out sideways. 
According to the present invention, when constituting the discharge guide 
portion in the flow guide portion of the partition wall for discharging 
the carry-over flow obliquely forwards relative to the advancing direction 
of the impeller, the position thereof and the angle of its surface 
positioned forward relative to the advancing direction of the impeller are 
particularly important. When the discharge guide portion is provided on 
the outer periphery side, it is desirable that an opening position on the 
side opposed to a vane of the impeller be on a more outer periphery side 
in the position opposed to the vane in a radial direction thereof, more 
preferably, that the opening position be outside a central part of the 
vane width in the radial direction, and still more preferably, it be on 
the outer periphery side 1/6 or more with respect to the central part of 
the vane in the radial direction of the vane. In the circumferential 
direction thereof, the opening position of the discharge guide portion on 
the side opposed to the vane is preferably determined so that a rear end 
of the flow guide portion of the partition wall is at a distance about 1.5 
to 2.5 times the vane-vane spacing with respect to a front end thereof in 
the advancing direction of the impeller. Further, the angle of the surface 
positioned forward relative to the impeller advancing direction, which is 
important for the jet of the carry-over flow, is preferably in the range 
from 5.degree. to 35.degree. relative to the impeller advancing direction, 
using as a reference plane the surface of the partition wall which defines 
a very small gap with respect to the impeller. 
When the discharge guide portion is provided on the inner periphery side, 
it is desirable that the opening position on the side opposed to the vane 
of the impeller be on a more inner periphery side in the position opposed 
to the vane, more preferably, that the opening position be inside a 
central part of the vane width in the radial direction, and still more 
preferably it be on the inner periphery side 1/6 or more with respect to 
the central part of the vane in the radial direction of the vane. In the 
circumferential direction thereof, the opening position of the discharge 
guide portion on the side opposed to the vane is preferably determined so 
that the rear end of the flow guide portion of the partition wall is at a 
distance about 1.5 to 2.5 times the vane-vane spacing with respect to the 
front end thereof in the advancing direction of the impeller. Further, the 
angle of the surface positioned forward relative to the impeller advancing 
direction, which is important for the jet of the carry-over flow, is 
preferably in the range from 5.degree. to 35.degree. relative to the 
impeller advancing direction, using as a reference plane the surface of 
the partition wall which defines a very small gap with respect to the 
impeller. 
By supplying an auxiliary flow to the fluid introduced from the suction 
port for conducting the fluid in the direction to form a rotating flow in 
the flow path, as mentioned above, the fluid which is apt to be disturbed 
near the suction port is dragged by the auxiliary flow in an enlarged 
state of the wetted perimeter length and is conducted in the rotating 
direction. Therefore, the entirety of the flow path can be used more 
effectively and the fluid in the flow path is rotated and pressurized by a 
larger number of vanes, whereby it is made possible to raise the pressure 
and improve the performance. Moreover, since the disturbance of fluid on 
the suction port side can be diminished by the auxiliary flow, it is 
possible to suppress noise. Further, in the case where a carry-over flow 
is utilized as the auxiliary flow, the carry-over flow which constitutes 
disturbance can be operated on rotation effectively, whereby a further 
improvement of the performance can be attained. 
These and further objects, features and advantages of the present invention 
will become more obvious from the following description when taken in 
connection with the accompanying drawings which show for purposes of 
illustration only, several embodiments in accordance with the present 
invention.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
Referring now to the drawings, FIGS. 1 and 5 illustrate a vortex flow 
blower such as a single-side impeller cup-type centrifugal pump having a 
plurality of vanes or vanes 1a in an annular groove 1b provided in an 
impeller 1. The impeller 1 is constructed to rotate on the center of a 
rotating shaft 14 driven by a prime mover 4. A casing 3 is provided with 
an inlet or suction port 3b and an outlet or discharge port 3c, and has a 
space for housing the impeller 1 inside. In the present embodiment, an 
induction motor is used as the prime mover 4. In a part of the casing 3 
facing the blades or vanes 1a of the impeller 1, there is provided an 
annular groove 3a of the casing (hereinafter referred to as the annular 
groove 3a) which is extends from the suction port 3b to the discharge port 
3c along the direction of rotation of the impeller 1, and which is open to 
the vanes 1a. 
The impeller 1 has a plurality of vanes 1a arranged so as to extend 
transversely to the annular groove 3a and the annular groove 1b of the 
impeller 1 (hereinafter referred to as the annular groove 1b of the 
impeller), which annular groove is disposed opposite to the annular groove 
3a across a small gap g as shown in FIG. 3(A). 
A part of the annular groove 3a on the circumference between the suction 
port 3b and the discharge port 3c is partitioned with a partition wall 
section 3d. The suction port 3b on the end side of rotation of the 
impeller 1 and the discharge port 3c on the start side of rotation of the 
impeller 1 are open at the bottom section of the annular groove 3a 
adjacent to the partition wall 3d. A sectional area reducer 3f is mounted 
for reducing a sectional area of the annular groove 3a in a part of zone 
extending from at least the discharge port 3c to a point midway o at the 
center (hereinafter referred to as the suction-discharge center 3e) of a 
portion of the annular groove 3a between the suction port 3b and the 
discharge port 3c as shown in FIG. 2. 
The area reducer, in the present embodiment, is a reduced section 3f as 
shown in FIGS. 2-4 and represented as a section cut on a plane passing a 
rotating shaft which is formed of a slanting projection extending from the 
vicinity of the outer peripheral edge of the annular groove 3a (the 
vicinity of the small gap g from the periphery of the impeller 1) to the 
bottom of this annular groove 3a. The area reducer is disposed in a zone 
extending from the suction-discharge center 3e of the annular groove 3a to 
the center of the discharge port 3c, and more particularly in a 
70-percent zone close to the suction-discharge center 3e, as shown in FIG. 
2. In the present invention, there exists an angle of 160.degree. between 
the suction-discharge center 3e and the discharge port 3c, and therefore 
the area reducer is disposed within a zone up to 112.degree. from the 
suction-discharge center 3e. The maximum range of zone in which the area 
reducer is provided starts at the suction-discharge center 3e, arriving at 
a position at the angle of 112.degree. along the annular groove in the 
direction of the discharge port, and the minimum range starts 
experimentally at a position at the angle of 30.degree. along the annular 
groove in the direction of the discharge port from the suction discharge 
center 3e and arrives at a position at an angle of 90.degree. from the 
suction discharge center 3e. The length of the zone of the area reducer 
corresponds to approximately 50 percent of the maximum value. 
The area reducer may be provided at the discharge or suction port section 
as described in the copending U.S. patent application Ser. No. 760,347. In 
this case, on the outlet or discharge side, a part of the internal flow of 
fluid hitting on the partition wall 3d is restrained to a smooth flow and 
furthermore enables noise reduction, whereas on the inlet side, the 
internal flow is guided to pass the vicinity of the impeller 1, being 
substantially accelerated by the impeller 1 to thereby increase the air 
volume. In the present embodiment, the prime mover 4 turns the impeller 1 
on the center of the rotating shaft 14 to produce an internal flow in the 
annular groove 3a and in the annular groove 1b of the impeller by the 
plurality of vanes 1a in the annular groove 1b of the impeller as shown in 
FIGS. 3(A) and 3(B). That is, the centrifugal pump of this invention is so 
constructed as to form the internal vortex flow of fluid whirling from the 
suction port 3b of the casing to the discharge port 3c through the 
suction-discharge center 3e and the area reducer 3f. 
In the present embodiment, there is formed the internal flow including a 
primary flow substantially accelerated in the direction of rotation of the 
impeller 1 from the suction port 3b to the middle 3c between the suction 
and discharge sides and a secondary flow whirling in the annular passages 
1b and 3a, the internal flow flowing smoothly without breakaway by flowing 
through the area reduced section 3f of a configuration exaggeratively 
shown by ridge lines in FIG. 4 from the middle 3e between the suction and 
discharge sides to the discharge port 3c. Thus, it is possible to prevent 
the occurrence of turbulence resulting from breakaway and accordingly to 
prevent the occurrence of noise and pressure loss. 
The internal flow in the vicinity of a no discharge operation where, for 
example, the discharge outlet is blocked, will become as shown in FIGS. 6 
to 8. In this case, the internal flow is rapidly guided to the inner 
periphery via the area reduced section 3f of the configuration 
exaggeratively shown by the ridge lines in FIG. 4 to thereby decrease the 
inflow velocity of the internal flow in the inner peripheral section of 
the impeller and to restrain noise occurrence and, at the same time, a 
loss likely to be caused by internal flow turbulence, thus obtaining an 
increased static pressure. The internal flow 30, as shown in FIG. 6, flows 
from the point S2 on the outer periphery of the impeller 1, flowing at a 
high rate into the annular groove 3a of the casing as far as the point S2 
in the direction of rotation of the impeller 1 on the outer periphery of 
the annular groove 3a. This flow does not flow to the discharge port 3c, 
but the fluid flows in a reverse direction of rotation on the inner 
periphery of the annular groove 3a and returns to the point S3 near the 
original outflow point S1 so that only an effective part of outflowing 
fluid in the annular groove 3a will flow. That is, the flow of the fluid 
on the outer and inner peripheries becomes as follows: 
##STR1## 
That is, the fluid flowing out at the point S1 on the outer periphery of 
the impeller 1 does not return to the point S1 when returning to the inner 
periphery of the impeller 1, but returns in the direction of rotation to 
the point S3 which has advanced by the carry-over flow rate QIK. 
If the passage in the annular groove 3a has a semicircular cross section, 
the fluid increased by a quantity corresponding to the angle of advance 
.theta.2 flows from the point S1 to the point S2 on the outer periphery of 
the annular groove 3a, and also the fluid increased by a quantity 
corresponding to the return angle of .theta.2' returns from the point S2 
back to the point S3 on the inner periphery of the annular groove 3a, and 
therefore the fluid from the point S3 flows into the impeller at a flow 
velocity w1 as shown in FIG. 8. Noise occurring in the no discharge 
operation of the centrifugal blower is largely attributable to a 
turbulence accompanying the inflow of fluid on the inner periphery of the 
impeller. A measured value of the internal flow indicates that the blower 
has the drawback that there occurs large noise and turbulence of flow when 
the fluid flows into the impeller because the flow velocity w1 is about 
twice as great as u2 which is the peripheral velocity of the impeller. 
Measurements of the internal flow of, for example, a centrifugal blower 
using an impeller of 210 mm diameter D2 and turning at a speed of 2850 rpm 
indicate that the peripheral velocity u2 of the impeller is 31.3 m/s, the 
flow velocity C2 at a no discharge operation is 78.5 m/s (C2 makes no 
difference between the presence and absence of the area reducer 3f in the 
casing), the flow velocity C1 is 6.5 m/s, and the flow velocity w1 is 93.5 
m/s, and further that, as shown in FIG. 10, a frequency component (fluid 
noise) at 200 to 1000 Hz and a frequency component (siren sound) at around 
2000 Hz are large, and an overall noise level is 63 db as obtained in a 
conventional centrifugal blower. According to the present embodiment, the 
flow can be changed from 30 to 30' as shown in FIG. 6, and the length of 
circular arcs S1 and S2 can be made shorter as compared with conventional 
ones by reducing the area of passage on the outer periphery of the annular 
groove 3a so as not to use the annular groove 3a of semicircular cross 
section as shown in FIG. 7(A). Therefore, the angle of advance of the 
fluid flow in the direction of rotation of the impeller also largely 
decreases from conventional .theta.2 to .theta.2'. 
With the decrease in the angle of advance of the fluid flow on the outer 
periphery of the annular groove 3a, the angle of return of fluid flow from 
the point S2 to the point S3 on the inner periphery of the annular groove 
3a also decreases. Therefore, the inflow velocity of fluid into the 
impeller becomes w1', much smaller than conventional w1, as shown in FIG. 
8. 
When the impeller diameter D2 is 210 mm, the inflow velocity w1' of fluid 
flowing into the impeller in the present embodiment becomes 65.2 mm, 
considerably smaller than the conventional w1 of 93.5 m/s. In consequence, 
a turbulence arising with the inflow of fluid flowing into the impeller 
largely decreases also, and the frequency component in the vicinity of 200 
to 1000 Hz and 2000 Hz attenuates, with a result that an overall noise 
level decreases as low as 56 db as shown in FIG. 9, which is 7 db lower 
than that provided by the construction of the vortex flow blower as 
described in the copending U.S. patent application Ser. No. 760,347. 
Furthermore, the power requirement for the blower are decreased by about 
20 percent or greater by lessening the turbulence. 
In the present embodiment, as described above, the annular groove in the 
casing of the cup-type centrifugal pump has, on its outer periphery, an 
area reducer for reducing a sectional area by a slant face which starts at 
the vicinity of a small gap on the outer periphery, and positively guides 
along the slant face the internal flow of fluid flowing out from the outer 
periphery of the impeller to the inner periphery, thereby preventing 
breakaway of flow in order to insure noise reduction and high static 
pressure. Therefore, the present invention has such advantages that the 
noise level can be lowered as low as about 7 dB, that is, a noise energy 
produced can be diminished to 20 percent, and, at the sane time, that the 
pressure can be increased by about 10 percent. Generally, the area reducer 
is formed as an integral part of the annular groove. However, the area 
reducer 3f may also be made of a member different from the casing and 
attached in the annular groove 3a, and may be so constituted as to guide 
the internal flow to the inner periphery along the slant face of this 
different member. The casing is generally produced of an aluminum die 
casting, but the area reducer 3f may be formed of a steel, ceramic, or 
fluoroplastic material. 
As shown in FIG. 3(A), for example, producing the area reducer 3f as a 
member different from the annular groove 3a can optimize the 
configuration, enabling the use of abrasion- and corrosion-resistant 
materials and also facilitating the replacement of the area reducer 3f. 
Therefore, it is possible to maintain the area reducer 3f in good 
condition even in a centrifugal pump which is exposed to twice as high an 
internal flow velocity as the peripheral velocity. 
Another embodiment of the present invention will be described with 
reference to FIGS. 11 to 15, wherein a position considered to be 
approximately at the center of the slant face of the area reducer 3f is 
specified, so that the function of the slant face will be more effectively 
performed. In the present embodiment, of a 70 percent part of the zone 
close to the suction-discharge center 3e, the zone extending from at least 
the suction-discharge center 3e of the annular groove 3a formed between 
the suction port 3b and the discharge port 3c to the center of the 
discharge port 3c, the area reducer 3f is constructed such that, as shown 
in FIG. 11, a point P1 beneath the passage surface of the area reducer 3f 
(the surface position of the slant) in the diameter of (3D2+D1)/4 in the 
annular groove 3a will be at a distance (D2-D1)/8 from the bottom face of 
the annular groove 3a in the center diameter (D2+D1)/2 in the annular 
groove 3a when no area reducer 3f is present in the annular groove. In the 
above description, D1 and D2 are the inner and outer peripheral diameters 
of the annular groove 3a, respectively, as measured from the center of 
the rotating shaft 14. In the present embodiment, there exists an angle of 
160.degree. between the suction-discharge center 3e to the center of the 
discharge port 3c. Therefore, the depth of the annular groove having the 
above-described relationship becomes shallow in the zone ranging from the 
suction-discharge center 3e to 112.degree.. This zone may extend to a 
70-percent part close to the suction-discharge center 3e of the zone 
extending from the suction-discharge center 3e to the center of the 
discharge port 3b as necessitated. 
On the discharge or outlet side from the suction-discharge center 3e, the 
components of the internal flow grow more in a circumferential direction 
than those in a direction of rotation. Therefore, a countermeasure for 
noise reduction is sufficient if performed mainly on the components of the 
internal flow in a circumferential direction. To reduce noise resulting 
from a turbulent flow, it is imperative to prevent breakaway of flow on 
the outer periphery. A conceivable method of preventing this breakaway is 
to decrease the sectional area of the annular groove, but when the 
sectional area is only decreased, the low passage will become too narrow, 
resulting in a decreased gas or air volume. 
In the present embodiment, the sectional area can be insured and 
accordingly a specific air volume can be maintained on the inner periphery 
by providing the area reducer 3f on the outer periphery as previously 
stated. It is possible to effectively prevent the air flow breakaway and 
to control a loss likely to be caused by an internal flow turbulence so as 
to increase the static pressure by providing the area reducer. 
FIG. 12 illustrates a variation of the present embodiment wherein the 
intermediate position of the area reducer 3f is shallower than the 
position P1 used as a reference position in FIG. 11, for more effective 
use of the area reducer 3f of the annular groove 3a. 
FIG. 13 illustrates another variation of the present embodiment wherein the 
area reducer 3f is provided in a part extending from the inner and outer 
peripheral edges of the annular groove 3a toward the bottom face of the 
annular groove. In this variation, the relationship with respect to the 
depth of the annular groove at the position P1 is the same as that shown 
in FIG. 11. 
FIG. 14 illustrates a further variation of the present embodiment wherein 
the area reducer 3f is formed of a slanting projection which starts, with 
a slight clearance provided, from the vicinity of a position (edge) with a 
small gap formed between the outer periphery of the impeller 1 and the 
outer periphery of the annular groove 3a, to the bottom face of the 
annular groove 3a. The area reducer 3f of the annular groove 3a is set 
with some clearance provided from the outer peripheral edge of the annular 
groove to leave some perpendicular section on the outer peripheral side, 
thus facilitating the positioning of the impeller. 
FIG. 15 illustrates another variation of the present embodiment wherein the 
outer peripheral position of the area reducer 3f of the annular groove 3a 
is set shallower than the position P1 used as a reference position in FIG. 
14, for the purpose of effective use of the area reducer 3f. In this 
variation, the area reducer 3f is so constructed that the depth, from the 
small gap face at D2, of a tangent between the point P1 in the diameter 
(3D2+D1)/4 of the annular groove 3a and a curve indicating the shape of a 
passage extending from the diameter (3D2+D1)/4 of the annular groove 3a in 
the same cross section to D2 will become less than (D2-D1)/10. 
A further embodiment of the centrifugal pump according to the present 
invention will be described with reference to FIG. 16, wherein the 
centrifugal pump also includes the impeller 1 and the casing 3 which has 
the suction port 3b and the discharge port 3c and houses the impeller 1 
therein. In a part facing the vanes 1a of the impeller in the casing 3, 
the annular groove 3a is formed along the direction of rotation of the 
impeller 1, extending from the suction port 3b to the discharge port 3c 
and opening to the vanes 1a. In this centrifugal pump, the area reducer 3f 
for reducing the sectional area in the annular groove 3a is provided in a 
part of the zone extending from at least the middle or midpoint 3e of a 
part of the annular groove 3a between the suction port 3b and the 
discharge port 3c and the center of the discharge port 3c. The depth of 
the annular groove 3a from the surface of the casing 3 facing the impeller 
1 increases in the order of an intermediate position between the middle 3e 
of the annular groove 3a and FIG. 16 shows cross sections of the annular 
groove at the positions A--A, C--C and D--D in the circumferential 
direction, in which order the depth of the annular groove increases. In 
the present embodiment, the use of the area reducer 3f can lower the noise 
level and obtain a high static pressure even within a range of large air 
volume and furthermore a specific air volume is obtainable because of a 
wide section D--D. 
Another embodiment of the present invention will be described with 
reference to FIGS. 17(A) and 17(B). In the present embodiment, the slant 
of the annular groove 3a is produced of the same member as the casing 3 
and the inner surface of the annular groove 3a is made in a slanting form 
protruding toward the inside of the annular groove within the range of 
slant formation of the casing, thereby providing the area reducer 3f. As 
shown in FIG. 17(A), the wall thickness T at the maximum thickness of the 
casing 3, within the range of slant formation, is two times larger than 
the wall thickness t of a part where no slant is formed in the same 
circumference. Using a casing with a thick-wall section as illustrated in 
FIG. 17(A) for forming the area reducer 3f can prevent occurrence of a 
problem if the casing becomes worn, and also enables high-rate manufacture 
of a quality device by using a mold cut to a desired form and also enables 
increasing durability by increasing the wall thickness of the casing. FIG. 
17(B) illustrates a variation of the present embodiment wherein the casing 
as a member of the annular groove 3a is formed into a slant face swelling 
toward the inside of the annular groove 3a without changing the plate 
thickness, thereby saving material and reducing the weight of the 
centrifugal pump. In the centrifugal pump, the maximum velocity of the 
internal flow is generally twice as high as the peripheral speed of the 
impeller. Because of such high velocity flow, the above-described 
consideration is needed. In the embodiments described above, the annular 
groove of a shape applicable to the single-side cup-type centrifugal pump 
has been described, but it is also possible to adapt the annular groove 
with a sectional area reducer to the double-side blade-type centrifugal 
pump. 
A further embodiment of the centrifugal pump, according to the present 
invention, will be described with reference to FIGS. 18 to 22 in relation 
to a double-side vane-type centrifugal pump. The basic construction of 
said centrifugal pump will be explained with reference to FIGS. 18 to 
20(A) and 20(B). The present embodiment of the double-side vane-type 
centrifugal pump provides for sectional area reduction by a slant face 
sloping toward the side of the impeller from the vicinity of a small gap, 
on the outer peripheral side of the annular groove in the casing, and 
positively guides the internal flow of fluid flowing out from the outer 
periphery of the impeller to the discharge port or to the inner periphery 
of the impeller along the slant face. 
The double-side vane-type centrifugal pump of the present embodiment 
comprises a double-side vane-type impeller 101 having on its outer 
periphery a number of vanes 101a protruding nearly radially in relation to 
a rotating shaft, a casing 103 having an annular groove 103a on the side 
and outer peripheral side correspondingly to the vanes 101a of the 
impeller 101 facing thereto, a side cover 115 having an annular groove 
115a which opens on the side and outer peripheral side correspondingly to 
the vanes 101a of the impeller 101 facing thereto, a partition wall 
section 103d which separates a part on the circumference of the annular 
groove 101a of the casing, a suction port 103b located adjacently to the 
partition wall section 103d of the casing and open in the axial side of 
the impeller 101, and a discharge port 103c located adjacent to the 
partition wall section 103d of the casing and open to the side facing to 
the rotating impeller. In this centrifugal pump, an area reducer 103f for 
reducing the sectional area in the annular groove 103a is provided in a 
part of a zone extending from at least the middle of the annular groove 
103a between the suction port 103b and the discharge port 103c to the 
discharge port 103c. 
In the present embodiment the double sided vane-type impeller 101 is driven 
by the prime mover, producing an internal flow as in the cup-type 
centrifugal pump. In this case, there is formed the internal flow 
consisting of a primary flow fully accelerated in the direction of 
rotation of the impeller 101 from the suction port 103b to the 
suction-discharge center 103e and a secondary flow whirling in vanes 101a 
and the annular passage 103a. The internal flow subsequently smoothly 
flows from the suction discharge center 103e to the discharge port 103c 
via the area reducer 103f without a breakaway of flow. 
FIGS. 19 and 20(A) and 20(B) show the basic operation of the double-side 
vane-type centrifugal pump of the present embodiment. The fluid flowing 
out from the center of the outer peripheral section of the impeller 101 is 
guided toward the annular groove 103a side through the area reducer 103f 
formed by a slanted portion. The internal flow flows out at a position 
slightly shifted from the center of the outer periphery of the impeller, 
being guided toward the annular groove 103a, close to the internal flow 
from the center of the outer periphery of the impeller. Therefore, there 
occurs little breakaway of flow occurs despite the difference in flow 
velocity between the fluid flowing out from the center of the outer 
periphery of the impeller in a conventional centrifugal pump and the fluid 
flowing out from a position shifted from the center of the outer periphery 
of the impeller, thereby obtaining a high static pressure by controlling 
the occurrence of sound and at the same time a loss resulting from the 
internal flow turbulence. According to the present embodiment, therefore, 
it is possible to lower the noise level and to increase the static 
pressure of the double-side blade-type centrifugal pump. 
The slant face as the area reducer is formed so that when D1 is the 
diameter at the gap face in the radial direction on the outer peripheral 
side of the annular groove 103a of the casing, D2 is the maximum diameter 
on the outer peripheral side of the annular groove 103a of the casing, g2 
is a side gap, and B is the width across faces, the depth of the passage 
surface in the diameter (D2+D1)/2 of the annular groove of the casing, in 
a 70-percent part close to the suction-discharge center of the zone 
extending from the suction-discharge center on the outer periphery of the 
annular groove of the casing to the center of the suction port and in a 
70-percent part close to the middle between the suction and discharge 
sides of the zone ranging from the suction-discharge center to the center 
of the discharge port will be over (D2-D1)/8 shallower than the maximum 
value of a radial depth on the outer peripheral side of the annular groove 
at the surface of the side gap g2. 
The area reducer 103f for reducing the sectional area of the annular groove 
by the use of a slant face may be a separate member attached as shown in 
FIG. 20(A) and 20(B), which guides the internal flow toward the impeller 
along the slant face thereof and to the discharge port or the inner 
periphery of the impeller. 
A variation of the present embodiment will be described with reference to 
FIG. 21 wherein the area reducer of the annular groove of the casing is 
formed of a slant face sloping sideward, starting at a position specified 
on the slant face, with some gap provided in the vicinity of the small gap 
g1, in order that the function of this slant face will be effectively 
effected. In this variation, as shown in FIG. 21, the area reducer 103f of 
the annular groove 13a is provided in the vicinity of, and with some 
clearance provided from, the gap g1 on the outer peripheral side of the 
annular groove 103a of the casing. This clearance is usable as a reference 
for positioning the impeller. 
Another variation of the present embodiment will be described with 
reference to FIG. 22 wherein the intermediate position of the area reducer 
103f of the annular groove 103a is set shallower than the position P1 used 
as a reference position in the first variation, for the purpose of 
effective use of the area reducer 103f. Further in this variation, as 
shown in FIG. 22, the depth from the small gap face g1 to P1 at the center 
of width on the outer periphery of the annular groove 103a of the casing 
is set so as to be (D2 - D1)/10 or less, to thereby decrease the sectional 
area of the annular groove 103a, thus improving the effect of leading the 
air flow into the inner peripheral side. 
As a further variation of the present embodiment, the depth of the annular 
groove from the small gap face of the casing on the outer peripheral side 
of the casing 103 may be increased (not illustrated) in the order of an 
intermediate position between the middle 103e between the suction and 
discharge sides and the suction port 103b of the casing, the middle 103e 
between the suction and discharge sides, and an intermediate position 
between the suction-discharge center 103e and the discharge port 103c. In 
this variation, on the suction side in the circumferential direction, no 
slant face is provided on the outer peripheral side of the annular groove 
of the casing, with a large air volume characteristics taken into 
consideration, so that the fluid flowing out of the outer periphery of the 
impeller will flow along the outer periphery, and that the flow will not 
cross the flow being drawn into the suction port. Further, according to 
this variation, the maximum air volume can be increased by about 20 
percent as compared with the volume of air flowing in the annular groove 
whose sectional area continues unchanged at a specific value from the 
suction port to the suction discharge center. Furthermore, in this 
variation, the thickness of the casing within the range of slant formation 
has been increased two times as large as the other part, thus increasing 
the thickness of the slant portion 103f to thereby prevent damage to the 
casing 103, such as a hole, resulting from abrasion by a high-velocity 
stream of fluid from the impeller 101 
According to the described embodiments of the present invention, the 
breakaway of the internal flow can be prevented by providing the annular 
groove sectional area reducer, thereby enabling lowering the noise level 
and increasing the static pressure throughout the range of air volume. 
A further embodiment of the present invention will now be described with 
reference to FIGS. 23 and 24 wherein FIG. 23 is a sectional view of a 
principal portion of the centrifugal pump showing the suction and 
discharge ports and the vicinity thereof and FIG. 24 is a sectional view 
taken along line 24--24. In these figures, the casing 3 has the annular 
groove 3a providing an annular flow path 208. The annular flow path 208 is 
in the form of the annular groove which is a generally semi-arcuate slot 
in its section which opens in a direction parallel to the axis of the 
rotating shaft 14 of the prime mover One end of the flow path 208 of the 
annular groove is in communication with a suction port 3b, while an 
opposite end thereof is in communication with a discharge port 3c. The 
section from the discharge port 3c to the suction port 3b is partitioned 
by a partition wall 3b which is opposed to the impeller through a very 
small gap. A suction-side passage 206 contiguous to the suction port 3b 
and a discharge-side passage contiguous to the discharge port 3c are 
provided in parallel within a silencer or muffler casing 5 which also 
serves as a base member as shown in FIG. 25. 
The impeller 1 is composed of a shroud and a plurality of blades or vanes 
and the shroud has an annular slot 211 which opens axially in an opposed 
relation to the annular flow path 208, centered on the rotating shaft 14. 
The opening portion of the annular flow path 208 and that of the shroud 
are opposed to each other by fixing the impeller 1 onto the rotating shaft 
14 of the prime mover or motor, whereby an annular flow path 212 having a 
circular section is formed. 
Upon rotation of the motor, the impeller 1 fixed onto the rotating shaft 14 
rotates. As a result, a fluid such as a gas which has been introduced from 
the suction port 3b rotates while describing a spiral or vortex flow as 
indicated by arrows in the annular flow path 212 of a circular section 
composed of the annular flow path 208 and the shroud under the action of 
the vanes 1a of the impeller, as shown in FIGS. 27 and 28. The gas is 
pressurized by the vanes 1a and is conveyed gradually in the rotating 
direction indicated at F. The thus-pressurized gas is conducted to the 
discharge port 6c by the action of the partition wall 3d and is discharged 
therefrom. 
As illustrated in FIG. 23, which is a sectional view of a principal portion 
showing a relation among the partition wall 3d, the suction and discharge 
ports 3b, 3c and the impeller 1, the partition wall 3d is provided at a 
front end portion thereof opposed to the impeller 1 with a flow guide 210 
such as plate member having a suction-side flow guide portion 210a for 
conducting the gas introduced from the suction port 3b smoothly to the 
annular flow path 212 and a discharge side flow guide portion 210b for 
conducting the gas which has been pressurized to the discharge port 3c 
smoothly. As is apparent also from this figure, the gas which has been 
pressurized by the impeller 1 is conducted to the discharge port 3c by the 
action of the partition wall 3d as indicated with arrow OUT. However, the 
outlet for gas 213 which was allowed to remain between adjacent vanes 1a 
at the time of discharge is closed with the partition wall 3d, so the gas 
213 is carried as it is to the suction port 3b side and is thus carried 
over to the suction side. This is called a carry-over flow. After decrease 
in the number of revolutions, the carry-over flow passes the partition 
wall 3d and is conveyed to the suction side. On the suction side, this 
pressurized fluid is released throughout the entire width of the vane 1a 
and expands without rotation substantially uniformly within the annular 
flow path 212. This expanded flow is mixed with gas introduced from the 
suction port 3b and indicated by arrow IN, thus disturbing the flow of the 
influent gas. Due to this disturbance, the gas which has been introduced 
through the suction port 3b from the exterior cannot start forming a 
rotating flow smoothly in the flow path portion near the suction port 3b, 
and only after passing this mixing region, it forms an effective rotating 
flow. According to an experimental measurement made by the present 
inventors, the mixing region reached 40.degree. in terms of the angle of 
circumference from the suction port 3b to the discharge port 3c side, as 
shown in FIG. 28. 
In this embodiment, in view of the point just mentioned above, a 
communication path 214 which is in communication with the suction port 3b 
from the surface side opposed to the vanes 1a is provided on the outer 
periphery side of the suction-side flow guide portion 210a of flow guide 
210 on the partition wall 3d, as shown in FIG. 23. This communication path 
constitutes an auxiliary flow supply path. The gas 213 remaining between 
adjacent vanes 1a passes through the communication path 214 before 
expanding in the vicinity of a front end of the suction-side flow guide 
portion 210a and is jetted to the suction port 3b side as indicated by 
arrow 215. The communication path 214 is provided at an angle which is 
obliquely forward relative to an advancing direction of the vanes 1a so 
that the gas jetted from the path 214 can rotate smoothly in the annular 
flow path 212. Of particular importance in this connection is a surface 
214a of the communication path 214 which surface in positioned forward 
relative to the advancing direction of the vanes 1a. The angle of the 
surface 214a in designated .alpha.. In this embodiment, the angle of a 
surface 214b positioned behind the surface 214a is also set at the same 
value. As to a radial position of the communication path 214 with respect 
to the impeller 1, the path 214 is disposed on a more outer periphery side 
of the vane 1a, as shown in FIG. 24. This is because the gas is compressed 
more outwards centrifugally by the vanes 1a and also because such position 
is advantageous to the formation of a rotating flow. 
According to this embodiment, since the communication path 214 is formed in 
the outer periphery portion of the suction-side flow guide portion 210a of 
the flow guide 210 on the partition wall 3d, the carry-over flow present 
on the outer periphery side of the impeller 1, of the gas 213 compressed 
and remaining between adjacent vanes 1a due to the presence of the 
partition wall 3d, flows out from the communication path 214 and forms a 
jet 215. The jet 215 flows to the inner periphery side of the casing 3 
along the inner wall of the casing, then further flows to the inner 
periphery side of the impeller 1 and forms a rotating flow. At this time, 
the gas present around the jet 215 is dragged by the jet and is conducted 
in the rotating direction. On the other hand, the compressed gas 213 which 
is allowed to remain between adjacent vanes 1a due to the presence of the 
partition wall 3d flows out from the outer periphery side under the action 
of the communication path 214, so that the inner periphery side between 
adjacent vanes 1a assumes a gas-free state and hence, in the vicinity of 
the suction port 3b immediately adjacent to the suction-side flow guide 
portion 210a, the gas which has been introduced from the exterior flows 
into the impeller easily. Thus, because of the generation of a rotating 
flow by the jet 215 and the easiness of the suction of gas to the inner 
periphery side of the impeller 1, a rotating flow 216 is formed smoothly 
in the vicinity of the suction port 3 b just after passing the 
suction-side flow guide portion 210a. At the same time, the wetted 
perimeter length also increases. Near the suction port 3b, therefore, the 
mixing region from the suction port 3b to the discharge port 3c becomes 
smaller and the wetted perimeter length increases, in comparison with the 
prior art, so that in this centrifugal blower the pressure rising action 
is enhanced in proportion to the angle of circumference and the wetted 
perimeter length. As a result, it becomes possible for the centrifugal 
blower to increase its discharge pressure and improve its performance. 
Further, since a rotating flow is formed smoothly in the vicinity of the 
suction port 3b just after passing the suction-side flow guide portion 
210a because of the generation of the rotating flow 216 by the jet 215 and 
because of the easiness of the suction of gas to the inner periphery side 
of the impeller 1, the disturbance of gas in this region decreases, so 
that the generation of noise can be much suppressed and there can be 
obtained a noise damping effect. 
According to an experimental measurement made by the present inventors, it 
was determined that an angle .alpha. of the communication path 214 
relative to the advancing direction of the vanes 1a in the range of 
5.degree. to 35.degree. was desirable in forming the rotating flow. In 
this embodiment, the angle .alpha. is set at 20.degree., and a radial size 
of the communication path 214 is set at a width of 1/3 of the vane width 
W. But this arrangement is for obtaining a greater effect. The radial size 
of the communication path 214 may cover the entire vane width, preferably 
on the outer periphery side. If a still greater effect is to be attained, 
it is desirable that the radial position of the communication path be on 
the outer side, more preferably 1/6 or more on the outer side, from the 
center of the vane width W. Further, as to a circumferential position of 
the communication path 214, a good result was obtained when a rear-end 
position B relative to the advancing direction of the vanes 1a was at a 
distance from a front-end position A of the discharge-side flow guide 
portion 210b in the range from 1.5 to 2.5 times the spacing between 
adjacent vanes 1a. However, the opening position of the communication path 
214 on the side opposed to the impeller 1 is not limited to such position 
if only the compressed gas remaining between adjacent vanes 1a can be 
introduced into the communication path. Not only the opening may be on the 
suction-side flow guide portion 210a as in the embodiment, but also it may 
span both the suction- and discharge-side flow guide portions 210a, 210b. 
FIG. 29 is a characteristic diagram showing air volume-static pressure 
characteristic of the centrifugal blower of this embodiment and that of a 
conventional centrifugal blower. In the same figure, a curve A represents 
an aerodynamic characteristic obtained in the presence of the 
communication path according to this embodiment, while a curve B 
represents an aerodynamic characteristic obtained in the absence of such 
communication path according to the prior art. These characteristics were 
obtained under the following conditions: motor used . . . 0.75 kW, 
effective dia. of the impeller . . . 235 mm, number of revolutions of the 
motor . . . 3,420 r.p.m., gap between the impeller and the partition wall 
. . . 0.3 mm, angle of the communication path . . . 20.degree.. As is 
apparent also from this figure, the aerodynamic characteristic in this 
embodiment could be improved about 20% as a whole in comparison with that 
in the prior art. 
According to another embodiment of the present invention, the invention is 
applied to an inner periphery side of a cup type centrifugal blower, such 
as a centrifugal gas pump. FIG. 30 is a sectional view of a principal 
portion thereof comprising suction and discharge ports and the vicinity 
thereof, and FIG. 31 is a sectional view taken along line 31--31 of FIG. 
30. FIG. 32 is a front view showing a state with a side cover and the 
impeller removed. 
The constructions of components, the principle of operation and problems 
involved in the conventional structure are the same as those referred to 
in the previous embodiment. 
According to this embodiment, as shown in FIG. 31, a communication path 214 
which is in communication with the suction port 3b side from its surface 
side opposed to a vane 1a, is provided on the inner periphery side of the 
suction-side flow guide portion 210a of the flow guide 210 provided on the 
partition wall 3d. This constitutes an auxiliary flow supply path. The gas 
213 remaining between adjacent vanes 1a passes through the communication 
path 214 before expanding in the vicinity of the front end of the 
suction-side flow guide portion 210a and is jetted to the suction port 3b 
side as indicated by arrow 215. The communication path 214 is provided at 
an angle of a obliquely forwards relative to the advancing direction of 
the vanes 1a so that the gas jetted from the path 214 can rotate smoothly 
in the annular flow path 212. Of particular importance in this connection 
is a surface 214a of the communication path 214 which surface is 
positioned forwards relative to the advancing direction of the vanes 1a. 
The angle of the surface 214a is designated .alpha.. As to a radial 
position of the communication path 214 with more inner periphery side of 
the vanes 1a, as shown in FIG. 31. This is for avoiding a delayed start of 
rotation on the inner periphery side while the gas is compressed more 
outwards centrifugally by the vanes 1a and rotation is started from the 
outer periphery side. 
According to this embodiment, since the communication path 214 is formed in 
the inner periphery portion of the suction-side flow guide portion 210a of 
the flow guide 210 on the partition wall 3d, the carry-over flow present 
on the inner periphery side of the impeller 1, of the gas 213 compressed 
and remaining between adjacent vanes 1a due to the presence of the 
partition wall 3d, flows out from the communication path 214 and forms a 
jet 215. The jet 215 flows to the inner periphery side of the casing along 
the inner wall of the casing, then further flows to the inner periphery 
side of the impeller 1 and forms a rotating flow. At this time, the gas 
present around the jet 215 is dragged by the jet and is conducted in the 
rotating direction. On the other hand, the compressed gas 213 which is 
allowed to remain between adjacent vanes 1a due to the presence of the 
partition wall 3d flows out from the communication path 214 and the 
pressure thereof is reduced. Consequently, on the inner periphery side 
between adjacent vanes 1a, and in the vicinity of the suction port 3b 
immediately after passing the suction side flow guide portion 210a of the 
flow guide 210 on the partition wall 3d, the gas introduced from the 
exterior easily flows into the impeller 1. Thus, because of the generation 
of a rotating flow by the jet 215 and the easiness of the suction of gas 
to the inner periphery side of the impeller 1, a rotating flow 216 is 
formed smoothly in the vicinity of the suction port 3a just after passing 
the suction-side flow guide portion 210a, and at the same time the wetted 
perimeter length also increases. Near the suction port 3b, therefore, the 
mixing region from the suction port to the discharge port 3c becomes 
smaller and the wetted perimeter length increase, in comparison with the 
prior art, so that in this centrifugal blower the pressure rising action 
is enhanced in proportion to the angle of circumference and the wetted 
perimeter length. As a result, it becomes possible for the centrifugal 
blower to increase its discharge pressure and improve its performance. 
Further, since a rotating flow is formed smoothly in the vicinity of the 
suction port 3b just after passing the suction-side flow guide portion 
210a because of the generation of the rotating flow 216 by the jet 215 and 
because of the easiness of the suction of gas to the inner periphery side 
of the impeller 1, the disturbance of gas in this region decreases, so 
that the generation of noise can be so much suppressed and there can be 
obtained a noise damping effect. 
According to an experimental measurement conducted by the present 
inventors, was determined out that an angle .alpha. of the communication 
path 214 relative to the advancing direction of the vanes 1a in the range 
of 5.degree. to 35.degree. was desirable in forming the rotating flow. In 
this embodiment, the angle .alpha. is set at 20.degree., and a radial size 
of the communication path 214 is set at a width of 1/3 of the vane width 
W. But this arrangement is for obtaining a greater effect. The radial size 
of the communication path 214 may cover the entire vane width, preferably 
on the inner periphery side. If a still greater effect is to be attained, 
it is desirable that the radial position of the communication path be 
about 1/3 or more on the inner side from the center of the vane width W. 
Further, as to a circumferential position of the communication path 214, a 
good result was obtained when a rear-end position B relative to the 
advancing direction of the vanes 1a was at a distance from a front-end 
position A of the discharge-side flow guide portion 210b in the range from 
1.5 to 2.5 times the spacing between adjacent vanes 1a. However, the 
opening position of the communication path 214 on the side opposed to the 
impeller 1 is not limited to such position if only the compressed gas 
remaining between adjacent vanes 1a can be introduced into the 
communication path. Not only the opening may be on the suction-side flow 
guide portion 210a as in this embodiment, but also it may span both the 
suction-side and discharge-side flow guide portions 210a, 210b. 
FIG. 33 is a characteristic diagram showing air volume - static pressure 
characteristic of the centrifugal blower of this embodiment and that of a 
conventional centrifugal blower. In the same figure, a curve A represents 
an aerodynamic characteristic obtained in the presence of the 
communication path according to this embodiment, while a curve B 
represents an aerodynamic characteristic obtained in the absence of such 
communication path according to the prior art. These characteristics were 
obtained under the following conditions: motor used . . . 0.75 kW, 
effective dia. of the impeller . . . 235 mm, number of revolutions of the 
motor . . . 3,420 r.p.m., gap between the impeller and the partition wall 
. . . 0.3 mm, angle of the communication path . . . 20.degree.. As is 
apparent also from this figure, the aerodynamic characteristic in this 
embodiment could be improved about 20% as a whole in comparison with that 
in the prior art. 
According to a further embodiment of the present invention, the invention 
is applied to an outer periphery side of a double-side vane type 
centrifugal blower, such as a centrifugal gas pump. FIG. 34 is a sectional 
view of a principal portion thereof comprising a suction port and the 
vicinity thereof, FIG. 35 is a sectional view taken along line 35--35 of 
FIG. 34, FIG. 36 is a sectional view taken along line 36--36 of FIG. 35, 
and FIG. 37 is a sectional side view showing the entire construction of 
this embodiment. 
In these figures, the numeral 1 denotes an impeller, numeral 3 denotes a 
casing which forms an annular flow path 208, and numeral 15 denotes a side 
cover which forms the annular flow path 208. The annular flow path 208 is 
in the form of a generally semi-arcuate slot in its section which opens in 
a direction parallel to the axis of a rotating shaft 14 of the prime 
mover. The flow path 208 is constituted in an annular shape, centered on 
the rotating shaft 14. One end of the flow path 208 is in communication 
with a suction port 3b, while an opposite end thereof is in communication 
with a discharge port 3c. The section from the discharge port 3c to the 
suction port 3b is partitioned by a partition wall 3d which is opposed to 
the impeller 1 through a very small gap. A suction-side passage 206 
contiguous to the suction port 3b and a discharge-side passage contiguous 
to the discharge port 3c are provided in parallel within a silencer or 
muffler casing 5 which also serves as a base member. 
The impeller 1 is composed a hub and a large number of blades or vanes 1a 
as shown in FIG. 36. The hub has an annular slot 211 which opens axially 
on both sides in an opposed relation to the annular flow path 208, 
centered on the rotating shaft 14. The vanes 1a are provided in a large 
number in a traversing direction for the slot 211. The opening portion of 
the annular flow path 208 and that of the hub are opposed to each other by 
fixing the impeller 1 onto the rotating shaft 14 of the primer mover, 
whereby an annular flow path 212 having a generally circular section is 
formed. 
Upon rotation of the prime mover, the impeller 1 fixed onto the rotating 
shaft 14 rotates. As a result, the gas which has been introduced from the 
suction port 3b rotates while describing a spiral flow as indicated by 
arrows in the annular flow path 212 of a circular section composed of the 
annular flow path 208 and the hub under the action of the vanes 1a of the 
impeller 1, as shown in FIGS. 35 and 36. The gas is pressurized by the 
vanes 1a and is conveyed gradually in the rotating direction. The 
thus-pressurized gas is conducted to the discharge port 3c by the action 
of the partition wall 3d and is discharged therefrom. 
As illustrated in FIG. 36, which is a sectional view of a principal portion 
showing a relation among the partition wall 3d, the suction and discharge 
ports 3b, 3c, and the impeller 1, the partition wall 3d is provided at a 
front end portion thereof opposed to the impeller 1 with a flow guide 210 
having a suction-side flow guide portion 210a for conducting the gas 
introduced from the suction port 3b smoothly to the annular flow path 212 
and a discharge-side flow guide portion 210b for conducting the gas which 
has been pressurized to the discharge port 3c. As is apparent also from 
this figure, the gas which has been pressurized by the impeller 1 is 
conducted to the discharge port 3c by the action of the partition wall 3d 
as indicated with arrow OUT. However, the outlet for gas 213 which was 
allowed to remain between adjacent vanes at the time of discharge is 
closed with the partition wall 3d, so the gas 213 is carried as it is to 
the suction port 3b side and is thus carried over to the suction side. 
This is called a carry-over flow. After decrease in the number of 
revolutions, the carry-over flow passes the partition wall 3d and is 
conveyed to the suction side. On the suction side, this pressurized fluid 
is released throughout the entire circumference of the vane 1a and expands 
without rotation substantially uniformly within the annular flow path 212. 
This expanded flow is mixed with gas introduced from the suction port 3b 
and indicated by arrow IN, thus disturbing the flow of the influent gas. 
Due to this disturbance, the gas which has been introduced through the 
suction port 3b from the exterior cannot start forming a rotating flow 
smoothly in the flow path portion near the suction port 3b, and only after 
passing this mixing region, it forms an effective rotating flow. 
In this embodiment, in view of the point just mentioned above, a 
communication path 214 which in communication with the suction port 3b 
from the surface side opposed to the vanes 1a, is provided on the outer 
periphery side of the suction-side flow guide portion 210a of the flow 
guide 210 on the partition wall 3d. This communication path constitutes an 
auxiliary flow supply path. The gas 213 remaining between adjacent vanes 
1a passes through the communication path 214 before expanding in the 
vicinity of a front end of the suction-side flow guide portion 210a and is 
jetted to the suction port 3b side as indicated by arrow 215. The 
communication path 214 is provided at an angle of .alpha. obliquely 
forwards relative to an advancing direction of the vanes 1a so that the 
gas jetted from the path 214 can rotate smoothly in the annular flow path 
212. Of particular importance in this connection is a surface 214a of the 
communication path 214 which surface is positioned forward relative to the 
advancing direction of the vanes 1a. The angle of the surface 214a is 
designated .alpha.. In this embodiment, the angle of a surface 214b 
positioned behind the surface 214a is also set at the same value. As to a 
radial position of the communication path 214 with respect to the impeller 
1, the path 214 is disposed on a more outer periphery side of the vane 19, 
as shown in FIG. 24. This is because the gas is compressed more outwards 
centrifugally by the vanes 1a and also because such position is 
advantageous to the formation of a rotating flow. 
According to this embodiment, since the communication path 214 is formed in 
the outer periphery portion of the suction-side flow guide portion 210a of 
the flow guide 210 on the partition wall 3d, the carry-over flow present 
on the outer periphery side of the impeller 1, of the gas 213 compressed 
and remaining between adjacent vanes 1a due to the presence of the 
partition wall 3d, flows out from the communication path 214 and forms a 
jet 215. The jet 215 flows to the inn.RTM.r periphery side of the casing 3 
along the inner wall of the casing, then further flows to the inner 
periphery side of the impeller 1 and forms a rotating flow. At this time, 
the gas present around the jet 215 is dragged by the jet and is conducted 
in the rotating direction. On the other hand, the compressed gas 213 which 
is allowed to remain between adjacent vanes 1a due to the presence of the 
partition wall 3d flows out from the outer periphery side under the action 
of the communication path 214, so that the inner periphery side between 
adjacent vanes 1a assumes a gas-free state and hence, in the vicinity of 
the suction port 3b immediately adjacent to the suction-side flow guide 
portion 210a, the gas which has been introduced from the exterior flows 
into the impeller easily. Thus, because of the generation of a rotating 
flow by the jet 215 and the easiness of the suction of gas to the inner 
periphery side of the impeller 1, a rotating flow 216 is formed smoothly 
in the vicinity of the suction port 3b just after passing the suction-side 
flow guide portion 210a. At the same time, the wetted perimeter length 
also increases. Near the suction port 3b, therefore, the mixing region 
from the suction port to the discharge port 3c becomes smaller and the 
wetted perimeter length increases, in comparison with the prior art, so 
that in this centrifugal blower the pressure rising action is enhanced in 
proportion to the angle of circumference and the wetted perimeter length. 
As a result, it becomes possible for the centrifugal blower to increase 
its discharge pressure and improve its performance. Further, since a 
rotating flow is formed smoothly in the vicinity of the suction port 3b 
just after passing the suction-side flow guide portion 210a because of the 
generation of the rotating flow 216 by the jet 215 and because of the 
easiness of the suction of gas to the inner periphery side of the impeller 
1, the disturbance of gas in this region decreases, so that the generation 
of noise can be so much suppressed and there can be obtained a noise 
damping effect. 
According to an experimental measurement made by the present inventors, it 
was determined that an angle of the communication path 214 relative to the 
advancing direction of the vanes 1a in the range of 5.degree. to 
35.degree. was desirable in forming the rotating flow. In this embodiment, 
the angle .alpha. is set at 12.degree., and a radial size of the 
communication path 214 is set at a width of 1/3 of the vane width W. But 
this arrangement is for obtaining a greater effect. The radial size of the 
communication path 214 may cover the entire vane width, preferably on the 
outer periphery side. If a still greater effect is to be attained, it is 
desirable that the radial position of the communication path be on the 
outer side, more preferably 1/6 or more on the outer side, from the center 
of the vane width W. Further, as to a circumferential position of the 
communication path 214, a good result was obtained when a rear-end 
position B relative to the advancing direction of the vanes 1a was at a 
distance from a front-end position A of the discharge-side flow guide 
portion 210b in the range from 1.5 to 2.5 times the spacing between 
adjacent vanes 1a. However, the opening position of the communication path 
214 on the side opposed to the impeller 1 is not limited to such position 
if only the compressed gas remaining between adjacent vanes 1a can be 
introduced into the communication path. Not only the opening may be on the 
suction-side flow guide portion 210a as in the embodiment, but also it may 
span both the suction-side and discharge-side flow guide portions 210a, 
210b. 
FIG. 38 is a characteristic diagram showing air volume - static pressure 
characteristic of the centrifugal blower of this embodiment and that of a 
conventional centrifugal blower. In the same figure, a curve A represents 
an aerodynamic characteristic obtained in the presence of the 
communication path according to this embodiment, while a curve B 
represents an aerodynamic characteristic obtained in the absence of such 
communication path according to the prior art. These characteristics were 
obtained under the following conditions: motor used . . . 0.75 kW, number 
of revolutions of the motor . . . 3,420 r.p.m., gap between the impeller 
and the partition wall . . . 0.3 mm, angle of the communication path . . . 
12.degree.. As is apparent also from this figure, the aerodynamic 
characteristic in this embodiment could be improved about 10% as a whole 
in comparison with that in the prior art. 
Although in the above-described embodiments, the gas which is allowed to 
remain between adjacent vanes 1a is utilized as the auxiliary flow, the 
gas present in another position may be utilized as the auxiliary flow, or 
gas which has been pressurized by another means may be utilized for the 
same purpose. Particularly in this case, an auxiliary flow supply path 
need not be provided in such a form as a communication hole in the 
partition wall 3d and hence the degree of freedom increases with respect 
to the position where such path is to be provided. However, as in the 
above embodiments, if the gas compressed and allowed to remain between 
adjacent vanes 1a by the partition wall 3d is utilized as the auxiliary 
flow, it is possible to utilize the gas which causes disturbance, and the 
wetted perimeter length increases, so there can be obtained outstanding 
effects in various points related to performance, including efficiency. 
Further, although a centrifugal blower has been described in each of the 
above embodiments, the present invention is not limited thereto. It goes 
without saying that the invention is applicable to centrifugal pumps in a 
broad sense, including centrifugal gas and liquid pumps. 
According to the present invention, as will be apparent from the above 
description, the performance of a centrifugal pump can be improved because 
it is possible to form a rotating flow more smoothly in an annular flow 
path. 
While we have shown and described several embodiments in accordance with 
the present invention, it is understood that the same is not limited 
thereto but is susceptible of numerous changes and modifications as known 
to those skilled in the art and we therefore do not wish to be limited to 
the details shown and described herein but intend to cover all such 
changes and modifications as are encompassed by the scope of the appended 
claims.