[TECHNICAL FIELD]
[0001] The present invention relates to a multiblade centrifugal fan.
[BACKGROUND ART]
[0002] There have been known multiblade centrifugal fans such as the sirocco fan, the turbo
fan and the radial fan. A multiblade centrifugal fan has numerous blades circumferentially
spaced from each other.
[0003] The noise of a multiblade centrifugal fan is raised by various factors such as the
fluid separation at the leading edges of the blades (inside edge of the impeller of
the fan) due to the difference between the fluid inlet angle to the leading edges
of the blades and the setting angle of the blades, the fluid separation in interblade
channels of the impeller of the fan, the difference between the fluid outlet angle
from the impeller of the fan and the divergence angle of a casing for accommodating
the impeller, the interference between the tongue of the casing and the blades, etc.
[0004] As shown in Figure 18(a), the fluid flows substantially radially to the leading edges
of the blades in an absolute coordinate system. However, as shown in Figure 18(b),
in a relative coordinate system seen from the blades, the fluid flows obliquely to
the direction in which the leading edges of the blades extend. In other words, the
fluid inlet angle to the leading edges of the blades is different from the setting
angle of the blades. Fluid separation at the leading edges of the blades is caused
as a result.
[0005] One counter measure for reducing the fluid separation at the leading edges of the
blades is, as is adopted in the turbo fan, to make the blades backward-curved to decrease
the difference between the fluid inlet angle to the leading edges of the blades and
the setting angle of the blades. However, the setting angle of the blades must be
greatly increased to achieve a substantial decrease in the noise of the fan, which,
in turn, degrades such hydrodynamic characteristics of the fan as the P-Q characteristics.
[DISCLOSURE OF INVENTION]
[0006] The object of the present invention is therefore to provide a multiblade centrifugal
fan wherein the difference between the fluid inlet angle to the leading edges of the
blades and the setting angle of the blades is reduced and the noise of the fan caused
by the fluid separation at the leading edges of the blades is reduced without changing
the setting angle of the blades.
[0007] According to the present invention, there is provided a multiblade centrifugal fan,
wherein numerous blades are disposed circumferentially spaced from each other, and
numerous annular plates are disposed radially inside the blades as stacked in the
direction in which the rotation axis of the multiblade centrifugal fan extends with
narrow intervening spaces between adjacent ones thereof.
[0008] In a multiblade centrifugal fan in accordance with the present invention, fluid flows
radially outward through channels formed between the numerous annular plates disposed
radially inside the blades as stacked in the direction in which the rotation axis
of the multiblade centrifugal fan extends with narrow intervening spaces between adjacent
ones thereof. The rotating annular plates apply tangential shear force to the fluid
flowing through the channels to accelerate it tangentially, thereby increasing its
tangential velocity. The fluid whose tangential velocity has been increased then flows
into channels formed between the adjacent blades. In the present multiblade centrifugal
fan, the difference between the circumferential velocity of the leading edges of the
blades and the tangential velocity of the fluid is smaller than that in a multiblade
centrifugal fan having no annular plates. Thus, in the present multiblade centrifugal
fan, the difference between the fluid inlet angle to the leading edges of the blades
and the setting angle of the blades is smaller than that in a multiblade centrifugal
fan having no annular plates.
[0009] According to a preferred embodiment of the present invention, the outer peripheries
of the annular plates are radially inwardly spaced from the leading edges of the blades.
[0010] Even if the outer peripheries of the annular plates are radially inwardly spaced
from the leading edges of the blades, the fluid flowing radially outward from the
channels between the adjacent annular plates flows to the leading edges of the blades
without losing tangential velocity. Thus, the difference between the fluid inlet angle
to the leading edges of the blades and the setting angle of the blades decreases.
The structure wherein the outer peripheries of the annular plates are radially inwardly
spaced from the leading edges of the blades is advantageous in that a multiblade centrifugal
fan in accordance with the present invention can be easily obtained by disposing the
annular plates in a conventional multiblade centrifugal fan.
[0011] According to another preferred embodiment of the present invention, the outer peripheries
of the annular plates are in contact with the leading edges of the blades.
[0012] According to another preferred embodiment of the present invention, the outer peripheries
of the annular plates overlap the leading edges of the blades.
[0013] When the outer peripheries of the annular plates are in contact with the leading
edges of the blades or overlap the leading edges of the blades, the fluid flowing
radially outward from the channels between the adjacent annular plates flows to the
leading edges of the blades without losing tangential velocity, thereby reducing the
difference between the fluid inlet angle to the leading edges of the blades and the
setting angle of the blades, and, in addition, the strength of the multiblade centrifugal
fan is increased by the contact or overlapping of the outer peripheries of the annular
plates with the leading edges of the blades.
[0014] According to another preferred embodiment of the present invention, the blades are
radially directed blades.
[0015] According to another preferred embodiment of the present invention, the blades are
backward-curved blades.
[0016] According to another preferred embodiment of the present invention, the blades are
forward-curved blades.
[0017] Irrespective of whether the blades are radially directed blades (radial fan), backward-curved
blades (turbo fan), or forward-curved blades (sirocco fan), the difference between
the fluid inlet angle to the leading edges of the blades and the setting angle of
the blades can be reduced by disposing numerous annular plates radially inside the
blades as stacked in the direction in which the rotation axis of the multiblade centrifugal
fan extends with narrow intervening spaces between adjacent ones thereof.
[BRIEF DESCRIPTION OF THE DRAWINGS]
[0018] In the drawings:
Figure 1 is a sectional view of a multiblade radial fan in accordance with a preferred
embodiment of the present invention.
Figure 2 is a sectional view taken along line II-II in Figure 1
Figure 3 is a sectional view of a multiblade radial fan showing the difference between
the fluid inlet angle to the leading edges of the blades and the setting angle of
the blades.
Figure 4 is a layout diagram of a measuring apparatus for measuring air volume flow
rate and static pressure of a multiblade centrifugal fan.
Figure 5 is a layout diagram of a measuring apparatus for measuring the sound pressure
level of a multiblade centrifugal fan.
Figure 6(a) is a plan view of a tested impeller (radial fan) without stacked annular
plates and Figure 6(b) is a sectional view taken along line b-b in Figure 6(a).
Figure 7(a) is a plan view of a tested impeller (radial fan) with stacked annular
plates and Figure 7(b) is a sectional view taken along line b-b in Figure 7(a).
Figure 8 is a plan view of a tested casing (radial fan).
Figure 9 is a view showing experimentally obtained correlations between minimum specific
sound levels KSmin and differences θ between the fluid inlet angle to the leading edges of the blades
and the setting angle of the blades.
Figure 10(a) is a sectional view of a tested impeller (sirocco fan) without stacked
annular plates and Figure 10(b) is a sectional view of a tested impeller (turbo fan)
without stacked annular plates.
Figure 11(a) is a sectional view of a tested impeller (sirocco fan) with stacked annular
plates and Figure 11(b) is a sectional view of a tested impeller (turbo fan) with
stacked annular plates.
Figure 12 is a plan view of a tested casing (sirocco fan).
Figure 13 is a plan view of a tested casing (turbo fan).
Figure 14 is a view showing a comparison between the noise level of a sirocco fan
with stacked annular plates and the noise level of a sirocco fan without stacked annular
plates (at an impeller rotation speed of 5100 rpm).
Figure 15 is a view showing a comparison between the noise level of a sirocco fan
with stacked annular plates and the noise level of a sirocco fan without stacked annular
plates (at an impeller rotation speed of 6120 rpm).
Figure 16 is a view showing a comparison between the noise level of a turbo fan with
stacked annular plates and the noise level of a turbo fan without stacked annular
plates (at an impeller rotation speed of 5100 rpm).
Figure 17 is a view showing a comparison between the noise level of a turbo fan with
stacked annular plates and the noise level of a turbo fan without stacked annular
plates (at an impeller rotation speed of 6120 rpm).
Figure 18(a) and Figure 18(b) are sectional views of a multiblade centrifugal fan
for xplaining why a difference arises between the fluid inlet angle to the leading
edges of the blades and the setting angle of the blades.
[THE BEST MODE FOR CARRYING OUT THE INVENTION]


A


1st embodiment
[0019] A multiblade radial fan in accordance with an embodiment of the present invention
will be described.
(I) Fan structure
[0020] In Figures 1 and 2, reference numeral 1 indicates a disk shaped base plate. An annular
top plate 2 is disposed above the base plate 1. The top plate 2 is disposed parallel
to and coaxially with the base plate 1. Numerous radial blades 3 are disposed as circumferentially
spaced from each other to connect the base plate 1 with the top plate 2. A plurality
of annular plates 4 are disposed radially inside the radial blades 3. The annular
plates 4 are disposed parallel to and coaxially with the base plate 1. The annular
plates 4 are stacked with narrow intervenig spaces between adjcent ones thereof. The
outer peripheries of the annular plates 4 fit tightly within horizontal slits formed
in the inner edges of the radial blades 3.
[0021] The base plate 1, the top plate 2, the radial blades 3 and the annular plates 4 constitute
an impeller 5. The central openings of the stacked annular plates 4 form a central
opening 5a of the impeller 5. Interplate channels 5b are formed between the base plate
1 and the lowermost annular plate 4, the top plate 2 and the uppermost annular plate
4, and adjacent annular plates 4. Interblade channels 5c are formed between adjacent
radial blades 3.
[0022] The impeller 5 is disposed in a casing 6 having a scroll shaped horizontal cross
section. The casing 6 is provided with an inlet opening 6a opposite the central opening
5a of the impeller 5 on the side of its top plate 2. The side wall of the casing 6
is provided with an outlet opening 6b and an outlet channel 7 is formed between the
outer periphery of the impeller 5 and the side wall of the casing 6.
[0023] A motor 8 is disposed below the casing 6. The motor 8 is fixed to the bottom plate
of the casing 6. The output shaft of the motor 8 extends upward through the bottom
plate of the casing 6 and is fixed to the center of the lower surface of the base
plate 1.
[0024] A multiblade radial fan having above described structure operates as follows.
[0025] The motor 8 starts. Fluid is drawn into the casing 6 through the inlet opening 6a.
The fluid drawn into the casing 6 flows into the interplate channels 5b. The fluid
entering the interplate channels 5b flows radially outward through the interplate
channels 5b. As indicated by double arrows in Figure 2, the base plate 1, the top
plate 2 and the annular plates 4, which are rotating, apply tangential shear force
to the fluid flowing through the channels 5b to accelerate it tangentially and apply
tangential velocity and centrifugal force to it. The fluid which has passed through
the interplate channels 5b flows into the interblade channels 5c. The fluid passing
into the interblade channels 5c flows radially outward through the interblade channels
5c. As indicated by single arrows in Figure 2, the radial blades 3, which are rotating,
apply force normal to the radial blades 3 to the fluid flowing through the channels
5c to accelerate it still more and apply still larger centrifugal force to it.
[0026] The fluid passing through the interblade channels 5c flows out of the outer ends
of the interblade channels 5c or the outer periphery of the impeller 5 and into the
outlet channel 7. The fluid flowing into the outlet channel 7 flows circumferentially
in the outlet channel 7 and flows out the casing 6 through the outlet opening 6b.
[0027] In the present multiblade radial fan, the base plate 1, the top plate 2 and the annular
plates 4 accelerate the fluid flowing through the interplate channels 5b tangentially
thereby increasing the tangential velocity thereof. Thus, in the present multiblade
radial fan, the difference between the circumferential velocity of the leading edges
of the radial blades 3 and the tangential velocity of the fluid flowing out of the
interplate channels 5b and into the interblade channels 5c is smaller than that in
a multiblade radial fan without the stacked annular plates 4. Thus, in the present
multiblade radial fan, the difference between the fluid inlet angle to the leading
edges of the blades 3 and the setting angle of the blades 3 is smaller than that in
a multiblade radial fan without the stacked annular plates 4, and the noise caused
by the fluid separation at the leading edges of the radial blades 3 is less than that
in a multiblade radial fan without the stacked annular plates 4. Outer peripheries
of the annular plates 4 fit tightly within horizontal slits formed in the inner edges
of the radial blades 3. Thus, the present multiblade radial fan is very sturdy.
(I I) Noise measurement
[0028] Noise measurements were carried out on multiblade radial fans in accordance with
the present invention and multiblade radial fans without the stacked annular plates.
(1) Difference between the inlet angle of the air to the leading edges of the blades
and the setting angle of the blades.
[0029] As shown in Figure 3, the radial direction of a multiblade radial fan is defined
as 0, the setting angle of the blades of the fan is defined as α , and the fluid inlet
angle to the leading edges of the blades is defined as β.
[0030] Then, the difference θ between the fluid inlet angle to the leading edges of the
blades of the multiblade radial fan and the setting angle of the blades of the multiblade
radial fan is given by formula ①.

In the above formula,
r
i : radial position of the leading edges of the blades
r₀ : radial position of the trailing edges of the blades
φ : flow coefficient

u₀ : mean radial flow velocity of the fluid at position r₀
c₀ : circumferential velocity of the blades at position r₀
[0031] The tangential velocity of the fluid flowing through the interplate channel relative
to the annular plates was obtained by the method of Hasinger (Hasinger, S. and Kehrt,
L., Trans. ASME, J.Eng.Power, 85(1963), 201). In accordance with this method, the
tangential velocity V
k of the fluid relative to the annular plates at the outer peripheries of the annular
plates is given by formula ②.


In the above formula,
- rj :
- inside radius of the annular plates
- rk :
- outside radius of the annular plates
- A :
- nondimensional constant

q : flow rate in an interplate channel
δ : space between adjacent annular plates
ν : kinematic viscosity
C
k : circumferential velocity of the annular plates at the position r
k
When stacked annular plates are disposed radially inside the leading edges of the
blades and the outer peripheries of the annular plates are in contact with the leading
edges of the blades, the difference angle θ at the leading edges of the blades is
given by formula ③ derived from the formulas ① and ②.

(2) Noise measurements
[0032] Noise measurements were carried out on multiblade radial fans in accordance with
the present invention and multiblade radial fans without the stacked annular plates
to obtain correlations between the minimum value of the specific sound level and the
difference angle θ.
〈1〉 Measuring apparatuses
① Measuring apparatus for measuring air volume flow rate and static pressure
[0033] The measuring apparatus used for measuring air volume flow rate and static pressure
is shown in Figure 4. The fan unit had an impeller 5, a scroll type casing 6 for accommodating
the impeller 5 and a motor 8. An inlet nozzle was disposed on the suction side of
the fan unit. A double chamber type air volume flow rate measuring apparatus (product
of Rika Seiki Co. Ltd., Type F-401) was disposed on the discharge side of the fan
unit. The air volume flow rate measuring apparatus was provided with an air volume
flow rate control damper and an auxiliary fan for controlling the static pressure
at the outlet of the fan unit. The air flow discharged from the fan unit was rectified
by a honeycomb.
[0034] The air volume flow rate of the fan unit was measured using orifices located in accordance
with the AMCA standard. The static pressure at the outlet of the fan unit was measured
through a static pressure measuring hole disposed near the outlet of the fan unit.
② Measuring apparatus for measuring sound pressure level
[0035] The measuring apparatus for measuring sound pressure level is shown in Figure 5.
An inlet nozzle was disposed on the suction side of the fan unit. A static pressure
control chamber of a size and shape similar to those of the air volume flow rate measuring
apparatus was disposed on the discharge side of the fan unit. The inside surface of
the static pressure control chamber was covered with sound absorbing material. The
static pressure control chamber was provided with an air volume flow rate control
damper for controlling the static pressure at the outlet of the fan unit.
[0036] The static pressure at the outlet of the fan unit was measured through a static pressure
measuring hole located near the outlet of the fan unit. The sound pressure level corresponding
to a certain level of the static pressure at the outlet of the fan unit was measured.
[0037] The motor 8 was installed in a soundproof box lined with sound absorbing material.
Thus, the noise generated by the motor 8 was confined.
[0038] The measurement of the sound pressure level was carried out in an anechoic room.
The A-weighted sound pressure level was measured at a point on the centerline of the
impeller and 1m above the upper surface of the casing.
〈2〉 Tested impellers, Tested casing
① Tested impellers without stacked annular plates
[0039] The outside diameter of the tested impellers (diameter at the trailing edges of the
radial blades 3) was fixed at 100 mm. The height of the tested impellers was fixed
at 24 mm. The thickness of the base plate 1 and the thickness of the top plate 2 were
both set at 2 mm. Three different impellers 5 without stacked annular plates 4 were
made. Different impellers 5 had a different ratio of the inside diameter (diameter
at leading edges of the radial blades 3) to the outside diameter, and a different
number of radial blades 3.
[0040] The particulars of the three tested impellers 5 (impeller numbers 1, 2 and 3) are
shown in Table 1, and Figures 6(a) and 6(b).
② Tested impellers with stacked annular plates
[0041] The outside diameter of the tested impellers (diameter at the trailing edges of the
radial blades 3) was fixed at 100 mm. The height of the tested impellers was fixed
at 24 mm. The thickness of the base plate 1 and the thickness of the top plate 2 were
both set at 2 mm. Three different impellers 5 with stacked annular plates 4 were made.
Different impellers 5 had a different ratio of the inside diameter (diameter at leading
edges of the radial blades 3) to the outside diameter, a different inside diameter
of the annular plate 4, and a different number of radial blades 3.
[0042] The particulars of the three tested impellers 5 (impeller numbers 4, 5 and 6) are
shown in Table 1, and Figures 7(a) and 7(b).
③ Tested casing
[0043] The height of the scroll type casing 6 was set at 27 mm. The divergence configuration
of the scroll type casing 6 was set as a logarithmic spiral defined by the following
formula. The divergence angle γ
c was set at 4.50°.

In the above formula,
r
c : radius of the side wall of the casing measured from the center of the impeller
5
r₀ : outside radius of the impeller 5
γ : angle measured from a base line, 0 ≦ γ ≦ 2π
γ
c : divergence angle
The tested casing 6 is shown in Figure 8.
③ Revolution speed of the impeller 5
[0044] The revolution speed of the impeller 5 was generally fixed at 6000 rpm but was varied
to a certain extent considering extrinsic factors such as background noise in the
anechoic room, condition of the measuring apparatus, etc. The revolution speeds of
the impellers 5 when the specific sound level became minimum are shown in Table 1.
(3) Measurement, Data Processing
〈2〉 Measurement
[0045] The air volume flow rate of the air discharged from the fan unit, the static pressure
at the outlet of the fan unit, and the sound pressure level were measured for each
of the 6 kinds of the impellers 5 shown in Table 1 when rotated at the revolution
speed shown in Table 1, while the air volume flow rate of the air discharged from
the fan unit was varied using the air volume flow rate control dampers.
(2) Data Processing
[0046] From the measured value of the air volume flow rate of the air discharged from the
fan unit, the static pressure at the outlet of the fan unit, and the sound pressure
level, a specific sound level K
S defined by the following formula was obtained.

In the above formula,
SPL(A) : A-weighted ( ∼ 20 KH
z ), 1/3 octave band overall sound pressure level, dB
Q : air volume flow rate of the air discharged from the fan unit, m³/s
P
t : total pressure at the outlet of the fan unit, mmAq
(4) Test Results
[0047] Based on the results of the measurements, a correlation between the specific sound
level K
S and the air volume flow rate was obtained for each tested impeller 5.
[0048] The correlation between the specific sound level K
S and the air volume flow rate Q was obtained on the assumption that a correlation
wherein the specific sound level K
S is K
S1 when the air volume flow rate Q is Q₁ exists between the specific sound level K
S and the air volume flow rate Q when the air volume flow rate Q and the static pressure
p at the outlet of the fan unit obtained by the air volume flow rate and static pressure
measurement are Q₁ and p₁ respectively, while the specific sound level K
S and the static pressure p at the outlet of the fan unit obtained by the sound pressure
level measurement are K
S1 and p₁ respectively The above assumption is thought to be reasonable as the size
and the shape of the air volume flow rate measuring apparatus used in the air volume
flow rate and static pressure measurement are substantially the same as those of the
static pressure controlling box used in the sound pressure level measurement.
[0049] The measurement showed that the specific sound level K
S of each tested impeller 5 varied with variation in the air volume flow rate. The
variation of the specific sound level K
S is caused by the casing 6. Thus, it can be assumed that the minimum value of the
specific sound level K
S or the minimum specific sound level K
Smin represents the noise characteristic of the tested impeller 5 itself free from the
effect of the casing 6, and the minimum specific sound level K
Smin does not include the sound level caused by the difference between the outlet angle
of the fluid flowing out the impeller and the divergence angle of the casing. In all
tested impellers, the relation between the ratio of the inside radius of the impeller
to the outside radius of the impeller and Karman-Millikan's nondimensional number
Z₁, which relation was referred to in the PCT application PCT/JP95/00789 filed by
the present applicant, was in the quite region which was proposed in the above PCT
application (Karman-Millikan's nondimensional numbers Z₁ are shown in Table 1). Thus,
it can be assumed that the minimum specific sound level K
Smin does not include the sound level caused by the fluid separation in the interblade
channels of the impeller. According to the result of a spectrum analysis of the measured
sound level data corresponding to the minimum specific sound level K
Smin , the energy of the sound with the same frequency as that of the noise due to the
interference between the tongue of the casing and the blades of the impeller was very
low. Thus, it can be assumed that the minimum specific sound level K
Smin does not include the sound level caused by the interference between the tongue of
the casing and the blades of the impeller.
[0050] From the above, it can be assumed that the minimum specific sound level K
Smin shows the noise characteristics of the impeller caused by the air separation at the
leading edges of the blades due to the difference between the air inlet angle to the
leading edges of the blades and the setting angle of the blades.
[0051] The minimum specific sound level K
Smin, the flow coefficient φ corresponding to the minimum specific sound levels K
Smin, and the difference angle θ corresponding to the minimum specific sound levels K
Smin of each tested impeller 5 are shown in Table 1. Correlations between the minimum
specific sound levels K
Smin and the difference angles θ of the tested impellers 5 are shown in Figure 9. The
difference angles θ were calculated on the assumption that the outside diameter of
the annular plate 4 (2 r
k ) is equal to the inside diameter of the impeller (the diameter at the leading edges
of the radial blades 3).
(5) Discussion
[0052] From Table 1 and Figure 9, it is clear that the difference angles θ of the impellers
5 with the stacked annular plates (impeller numbers 4, 5 and 6) are smaller than those
of the impellers 5 without the stacked annular plates (impeller numbers 1, 2 and 3),
and the minimum specific sound level K
Smin decreases as the difference angle θ decreases.
[0053] From the above described sound level measurement, it was confirmed that the present
invention can effectively reduce the noise of a multiblade centrifugal fan caused
by the fluid separation at the leading edges of the blades due to the difference between
the fluid inlet angle to the leading edges of the blades and the setting angle of
the blades.


B


2nd embodiment
(1) Noise measurement
[0054] Noise measurements were carried out on sirocco fans and turbo fans produced by Rokugo
Seisakusho Co. Ltd. The noise measurements were carried out on fans with stacked annular
plates and fans without stacked annular plates. From the noise measurements, it was
confirmed that the present invention is also effective when applied to sirocco fans
and turbo fans.
〈1〉 Measuring apparatuses
[0055] The same measuring apparatuses for measuring air volume flow rate and static pressure
as those used in the 1st embodiment were used. The same measuring apparatuses for
measuring sound pressure level as those used in the 1st embodiment were used.
〈2〉 Tested impellers, Tested casing
① Tested impellers without stacked annular plates
[0056] An impeller of a sirocco fan produced by Rokugo Seisakusho Co. Ltd., (impeller no.1
) ( outside diameter × inside diameter × height of the interblade channels × number
of blades = 102.0 mm × 85.3 mm × 29.0 mm × 32 ), and an impeller of a turbo fan produced
by Rokugo Seisakusho Co. Ltd., (impeller no.2 ) ( outside diameter × inside diameter
× height of the interblade channels × number of blades = 99.0 mm × 54.8 mm × 17.0
mm × 10 ) were used as tested impellers.
[0057] The particulars of the no. 1 impeller are shown in Table 2 and Figure 10(a). The
particulars of the no. 2 impeller are shown in Table 2 and Figure 10(b).
② Tested impellers with stacked annular plates
[0058] An impeller constituted by providing the impeller no. 1 with stacked annular plates
(outside diameter × inside diameter × thickness of the annular plates × number of
the annular plates × space between the adjacent annular plates × height = 85.0 mm
× 65.0 mm × 0.3 mm × 42 × 0.4 mm × 29.0 mm) (impeller no. 3), and an impeller constituted
by providing the impeller no. 2 with stacked annular plates (outside diameter × inside
diameter × thickness of the annular plates × number of the annular plates × space
between the adjacent annular plates × height = 54.0 mm × 40.0 mm × 0.3 mm × 22 × 0.4
mm × 15.0 mm) (impeller no. 4) were used as tested impellers.
[0059] The particulars of the no. 3 impeller are shown in Table 2 and Figure 11(a). The
particulars of the no. 4 impeller are shown in Table 2 and Figure 11(b).
③ Tested casing
[0060] The height of the scroll type casing was set at an impeller height (interblade channel
height + base plate thickness + top plate thickness) of + 9 mm for the impeller of
the sirocco fan, and an impeller height (interblade channel height + base plate thickness
+ top plate thickness) of + 8 mm for the impeller of the turbo fan. The divergence
configuration of the scroll type casing was set as a logarithmic spiral defined by
the following formula. The divergence angle γ
c was set at 4.50°.

In the above formula,
r
c : radius of the side wall of the casing measured from the center of the impeller
5
r₀ : outside radius of the impeller 5
γ : angle measured from a base line, 0 ≦ γ ≦ 2π
γ
c : divergence angle
The tested casing for the impellers no. 1 and no. 3 (sirocco fan) is shown in Figure
12. The tested casing for the impellers no. 2 and no. 4 (turbo fan) is shown in Figure
13.
④ Revolution speed of the impeller
[0061] The revolution speed of the impeller was set at 5100 rpm and 6120 rpm.
(2) Measurement, Data Processing
〈1〉 Measurement
[0062] The air volume flow rate of the air discharged from the fan unit, the static pressure
at the outlet of the fan unit, and the sound pressure level were measured for each
of the 4 kinds of the impellers shown in Table 2 when rotated at the aforesaid two
revolution speeds, while the air volume flow rate of the air discharged from the fan
unit was varied using the air volume flow rate control dampers.
(2) Data Processing
[0063] Specific sound levels K
S were obtained in the same way as in the 1st embodiment.
(3) Test Results
[0064] Based on the results of the measurements, a correlation between the specific sound
level K
S and the air volume flow rate Q was obtained for each tested impeller in the same
way as in the 1st embodiment. Flow coefficients φ were obtained from the air volume
flow rates Q based on the following formula ④.

In the above formula,
u = Q / S : radial velocity of the air flow at the outlet of the impeller
v = r ω : circumferential velocity of the outer periphery of the impeller
S = 2 πrh : area of the outlet of the impeller
Q : air volume flow rate
r : outside radius of the impeller
h : interblade channel height of the impeller
ω : rotation speed of the impeller
[0065] Correlations between the specific sound levels K
S and the flow coefficients φ were obtained from the correlations between the specific
sound levels K
S and the air volume flow rates Q and the flow coefficients φ derived from the air
volume flow rates Q.
[0066] Correlations between the specific sound levels K
S and the flow coefficients φ are shown in Figures 14 to 17.
[0067] From Figures 14 to 17, it is clear that also in the sirocco fan and the turbo fan
the fan noise can be reduced over a wide range of flow coefficient distribution by
disposing the stacked annular plates in the impeller.
[0068] It seems that the above described noise reduction was achieved by the combined effect
of the noise reduction due to the suppression of the fluid separation at the leading
edges of the blades, the noise reduction due to the suppression of the fluid separation
in the interblade channels following the suppression of the fluid separation at the
leading edges of the blades, and the reduction of the interference noise between the
tongue of the casing and the blades resulting from more uniform circumferential velocity
distribution of the air flow at the outlet of the interblade channels following the
suppression of the fluid separation at the leading edges of the blades and in the
interblade channels.
[0069] As described above, it is clear that also in the sirocco fan and the turbo fan the
fan noise caused by the fluid separation at the leading edges of the blades can be
reduced by disposing the stacked annular plates in the impeller.
(4) Discussion
[0070] It was confirmed that the present invention is effective when applied to the sirocco
fan and the turbo fan.
[0071] Although preferred embodiments of the present invention and the noise measurements
for confirmation of the effectiveness of the present invention were described above,
the present invention is not restricted to the above mentioned embodiments.
[0072] For example, the outer peripheries of the annular plates 4 of the multiblade radial
fan of the 1st embodiment may be radially inwardly spaced from the leading edges of
the radial blades 3 or be in contact with the leading edges of the radial blades.
[0073] Even if the outer peripheries of the annular plates 4 are radially inwardly spaced
from the leading edges of the blades 3, the fluid flowing radially outward from the
channels between the adjacent annular plates 4 flows to the leading edges of the blades
3 without losing tangential velocity. Thus, the difference between the fluid inlet
angle to the leading edges of the blades 3 and the setting angle of the blades 3 decreases.
The structure wherein the outer peripheries of the annular plates 4 are radially inwardly
spaced from the leading edges of the blades 3 is advantageous in that a multiblade
radial fan in accordance with the present invention can be easily obtained by disposing
the annular plates in a conventional multiblade radial fan.
[0074] When the outer peripheries of the annular plates 4 are in contact with the leading
edges of the blades 3, the same effects can be achieved as when the outer peripheries
of the annular plates 4 overlap the leading edges of the blades 3. Specifically, the
fluid flowing radially outward from the channels between the adjacent annular plates
4 flows to the leading edges of the blades 3 without losing tangential velocity, thereby
reducing the difference between the fluid inlet angle to the leading edges of the
blades 3 and the setting angle of the blades 3, and, in addition, the strength of
the multiblade radial fan can be increased by brazing, bonding or otherwise fixing
the outer peripheries of the annular plates 4 in contact with the leading edges of
the blades 3.
[0075] In the multiblade radial fan of the 1st embodiment, the stacked annular plates 4
need not be disposed over the entire space between the base plate 1 and the top plate
2 but instead may be disposed only over the portion of the space near the base plate
1 or over the portion of the space near the top plate 2 or over the mid-portion of
the space.
[INDUSTRIAL APPLICABILITY OF THE INVENTION]
[0076] The present invention provides a multiblade centrifugal fan wherein the difference
between the fluid inlet angle to the leading edges of the blades and the setting angle
of the blades is reduced and the noise of the fan caused by the fluid separation at
the leading edges of the blades is reduced without changing the setting angle of the
blades.
TABLE 2
impeller no. |
1 |
2 |
3 |
4 |
outside diameter (mm) |
102.0 |
99.0 |
102.0 |
99.0 |
inside diameter (mm) |
85.3 |
54.8 |
85.3 |
54.8 |
height of interblade channel (mm) |
29.0 |
17.0 |
29.0 |
17.0 |
number of blades |
32 |
10 |
32 |
10 |
outside diameter of annular plate (mm) |
|
|
85.0 |
54.0 |
inside diameter of annular plate (mm) |
|
|
65.0 |
40.0 |
thickness of annular plate (mm) |
|
|
0.3 |
0.3 |
number of annular plates |
|
|
42 |
22 |
space between adjacent annular plates (mm) |
|
|
0.4 |
0.4 |
height of stacked annular plates (mm) |
|
|
29.0 |
15.0 |