EP0717194A1 - Multivane centrifugal fan - Google Patents

Multivane centrifugal fan Download PDF

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Publication number
EP0717194A1
EP0717194A1 EP95923554A EP95923554A EP0717194A1 EP 0717194 A1 EP0717194 A1 EP 0717194A1 EP 95923554 A EP95923554 A EP 95923554A EP 95923554 A EP95923554 A EP 95923554A EP 0717194 A1 EP0717194 A1 EP 0717194A1
Authority
EP
European Patent Office
Prior art keywords
blades
annular plates
fan
impeller
leading edges
Prior art date
Legal status (The legal status is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the status listed.)
Withdrawn
Application number
EP95923554A
Other languages
German (de)
English (en)
French (fr)
Inventor
Makoto Hatakeyama
Noboru Shinbara
Hisato Haraga
Current Assignee (The listed assignees may be inaccurate. Google has not performed a legal analysis and makes no representation or warranty as to the accuracy of the list.)
Toto Ltd
Original Assignee
Toto Ltd
Priority date (The priority date is an assumption and is not a legal conclusion. Google has not performed a legal analysis and makes no representation as to the accuracy of the date listed.)
Filing date
Publication date
Application filed by Toto Ltd filed Critical Toto Ltd
Publication of EP0717194A1 publication Critical patent/EP0717194A1/en
Withdrawn legal-status Critical Current

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Classifications

    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D29/00—Details, component parts, or accessories
    • F04D29/40—Casings; Connections of working fluid
    • F—MECHANICAL ENGINEERING; LIGHTING; HEATING; WEAPONS; BLASTING
    • F04—POSITIVE - DISPLACEMENT MACHINES FOR LIQUIDS; PUMPS FOR LIQUIDS OR ELASTIC FLUIDS
    • F04D—NON-POSITIVE-DISPLACEMENT PUMPS
    • F04D17/00—Radial-flow pumps, e.g. centrifugal pumps; Helico-centrifugal pumps
    • F04D17/08—Centrifugal pumps
    • F04D17/16—Centrifugal pumps for displacing without appreciable compression
    • F04D17/161—Shear force pumps

Definitions

  • the present invention relates to a multiblade centrifugal fan.
  • 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.
  • 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.
  • the fluid flows substantially radially to the leading edges of the blades in an absolute coordinate system.
  • the fluid flows obliquely to the direction in which the leading edges of the blades extend.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • the outer peripheries of the annular plates are radially inwardly spaced from the leading edges of the blades.
  • 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.
  • the outer peripheries of the annular plates are in contact with the leading edges of the blades.
  • the outer peripheries of the annular plates overlap the leading edges of the blades.
  • the blades are radially directed blades.
  • the blades are backward-curved blades.
  • the blades are forward-curved blades.
  • 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.
  • a multiblade radial fan in accordance with an embodiment of the present invention will be described.
  • 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.
  • 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.
  • 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.
  • 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.
  • a multiblade radial fan having above described structure operates as follows.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • 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.
  • the present multiblade radial fan is very sturdy.
  • Noise measurements were carried out on multiblade radial fans in accordance with the present invention and multiblade radial fans without the stacked annular plates.
  • the radial direction of a multiblade radial fan is defined as 0, the setting angle of the blades of the fan is defined as ⁇ , and the fluid inlet angle to the leading edges of the blades is defined as ⁇ .
  • 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 ⁇ .
  • 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.
  • 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.
  • 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.
  • 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.
  • the motor 8 was installed in a soundproof box lined with sound absorbing material. Thus, the noise generated by the motor 8 was confined.
  • 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.
  • 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.
  • 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.
  • 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 ⁇ c was set at 4.50°.
  • r c r0 exp ( ⁇ tan ⁇ c )
  • 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
  • ⁇ angle measured from a base line
  • 0 ⁇ ⁇ ⁇ 2 ⁇ ⁇ c divergence angle
  • the tested casing 6 is shown in Figure 8.
  • 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.
  • 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.
  • K S SPL(A)-10log 10 Q( P t ) 2
  • 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
  • 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.
  • the variation of the specific sound level K S is caused by the casing 6.
  • 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.
  • 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.
  • 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.
  • the minimum specific sound level K Smin , the flow coefficient ⁇ corresponding to the minimum specific sound levels K Smin , and the difference angle ⁇ 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 ⁇ of the tested impellers 5 are shown in Figure 9. The difference angles ⁇ 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).
  • 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.
  • 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.
  • 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 ⁇ c was set at 4.50°.
  • the revolution speed of the impeller was set at 5100 rpm and 6120 rpm.
  • 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.
  • Correlations between the specific sound levels K S and the flow coefficients ⁇ were obtained from the correlations between the specific sound levels K S and the air volume flow rates Q and the flow coefficients ⁇ derived from the air volume flow rates Q.
  • 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.
  • the same effects can be achieved as when the outer peripheries of the annular plates 4 overlap 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, 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.
  • 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.
  • 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.

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  • Engineering & Computer Science (AREA)
  • Mechanical Engineering (AREA)
  • General Engineering & Computer Science (AREA)
  • Structures Of Non-Positive Displacement Pumps (AREA)
EP95923554A 1994-06-30 1995-06-30 Multivane centrifugal fan Withdrawn EP0717194A1 (en)

Applications Claiming Priority (3)

Application Number Priority Date Filing Date Title
JP17048494 1994-06-30
JP170484/94 1994-06-30
PCT/JP1995/001307 WO1996000854A1 (fr) 1994-06-30 1995-06-30 Ventilateur centrifuge a aubes multiples

Publications (1)

Publication Number Publication Date
EP0717194A1 true EP0717194A1 (en) 1996-06-19

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EP95923554A Withdrawn EP0717194A1 (en) 1994-06-30 1995-06-30 Multivane centrifugal fan

Country Status (5)

Country Link
EP (1) EP0717194A1 (enrdf_load_stackoverflow)
KR (1) KR960704162A (enrdf_load_stackoverflow)
CA (1) CA2168551A1 (enrdf_load_stackoverflow)
TW (1) TW268078B (enrdf_load_stackoverflow)
WO (1) WO1996000854A1 (enrdf_load_stackoverflow)

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
SE2350930A1 (en) * 2023-07-27 2025-01-28 Swegon Operations Ab An array of radial fan arrangements

Families Citing this family (2)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
ITBO20050432A1 (it) 2005-06-29 2006-12-30 Giorgio Pecci Procedimento di trasformazione di sostanze organiche solide aventi legami chimici in catene lunghe in miscugli di componenti solidi e/o liquidi e/o gassosi organici e/o inorganici a catene corte ed apparecchiatura per attuarlo
TWI457505B (zh) * 2009-01-08 2014-10-21 Delta Electronics Inc 通風裝置及其葉輪

Family Cites Families (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
JPS5666494A (en) * 1979-11-02 1981-06-04 Hitachi Ltd Impeller for centrifugal compressor

Non-Patent Citations (1)

* Cited by examiner, † Cited by third party
Title
See references of WO9600854A1 *

Cited By (1)

* Cited by examiner, † Cited by third party
Publication number Priority date Publication date Assignee Title
SE2350930A1 (en) * 2023-07-27 2025-01-28 Swegon Operations Ab An array of radial fan arrangements

Also Published As

Publication number Publication date
KR960704162A (ko) 1996-08-31
TW268078B (enrdf_load_stackoverflow) 1996-01-11
WO1996000854A1 (fr) 1996-01-11
CA2168551A1 (en) 1996-01-11

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