WO1998045601A1

WO1998045601A1 - Centrifugal fan with flow control vanes

Info

Publication number: WO1998045601A1
Application number: PCT/US1998/005034
Authority: WO
Inventors: Martin G. Yapp
Original assignee: Bosch Automotive Systems Corporation
Priority date: 1997-04-04
Filing date: 1998-03-13
Publication date: 1998-10-15
Also published as: AU6465398A

Abstract

A centrifugal fan (10) and methods of using such a fan are provided including a rotating impeller (12) comprising a hub (22), a rotating shroud (24) and blades (18) extending axially from the shroud to the hub, in a casing (14) which has an inlet shroud (26) defining an inlet for incoming flow, there being an annular space (28) between the rotating shroud and the inlet shroud. The inlet shroud has at least one cambered flow control vane (46) having a flow surface arranged to redirect recirculating flow from the annular space. The flow surface extends from a radially outer edge to a radially inner edge and has a tangential direction component opposite to the fan rotation direction at the radially inner edge, such that the recirculating flow through the annular space is redirected to give it a tangential velocity component opposite the fan rotation direction, which improves fan output pressure, efficiency, and noise.

Description

Centrifugal Fan with Flow Control Vanes

Background of the Invention This invention is in the general field of centrifugal fans or blowers.

Centrifugal -flow bladed impellers mass-produced from injection molded thermoplastics for use in combination with a housing or shroud in a blower typically operate with substantial running clearance between the moving impeller and the stationary housing or shroud. A common running clearance is, for example, 0.015 to 0.04 times the diameter of the fan. The pressure rise across centrifugal fans is considerable, and drives a significant air flow through the running clearance gap between the shroud and impeller. This typically undesirable running clearance flow is sometimes called "recirculation flow", and has been measured to be, in some instances, 14 to 27 percent of the flow through the fan. We use the term fan or blower to describe centrifugal-flow impellers in combination with a housing (e.g. a shroud or a duct) that guides airflow entering and/or leaving the fan. The impeller rotates in a direction on an axis and has a rotating hub and rotating shroud, with blades extending generally axially from the shroud to the hub. Examples of such fans have commercial use in automotive and residential cooling and ventilation applications .

One way to reduce recirculation flow is to reduce the clearance gap between the moving impeller and the stationary shroud. Another way to reduce recirculation flow is to incorporate labyrinth seals or baffles between the impeller and the shroud (see, e.g., U.S. Pat. Nos . 3,782,851 and 4,357,914). Recirculation flow tends to have a swirl component, an angular velocity in the direction of the rotation of the impeller. Because this swirl can cause reductions in fan pressure, sometimes accompanied by an increase in audible noise, various attempts have been made to reduce the angular velocity of the recirculation flow and to direct it radially inward. McDonald U.S. Pat. No. 3,285,501 and White U.S. Pat. No. 3,307,776 each disclose an example of stationary flow control vanes designed to decrease or remove such swirl or spin and stabilize the recirculation flow.

Summary of the Invention It has been realized that a potential drawback of the use of flow control vanes or "leakage stators" is undesirable noise caused by jet-wake flows produced by each vane. Under some operating conditions a flow structure, or pattern of velocity variation, corresponding to the vane spacing is established in the recirculation flow and ingested into the fan, causing undesirable increases in noise. These noise increases can be characterized by the addition of tonal energy at blade passing frequency and its harmonics, as well as increases in broadband noise at higher frequencies due to the turbulent nature of the vane-produced jets. It has been found that an improved configuration of leakage stator flow control vanes and arrangement of the recirculation flow path can result in advantages of increased output pressure, higher static efficiency (i.e. fluid power per shaft power) and/or improved noise characteristics.

According to one aspect of the invention, a centrifugal fan has a rotating impeller and a casing with an inlet shroud with a flow control vane arranged to redirect the recirculating airflow so as to substantially increase its velocity component in a direction tangential to the impeller shroud and opposite to the rotational direction of the impeller. The inlet shroud defines an inlet for incoming flow, and there is an annular space between the rotating shroud and the inlet shroud for running clearance, such that recirculating airflow moves through the annular space toward the inlet.

In another aspect, an annular space is defined between the rotating shroud and the inlet shroud for running clearance, and the inlet shroud has a cambered flow control vane having a flow surface arranged to redirect recirculating flow from the annular space. The flow surface extends from a radially outer edge to a radially inner edge, and the projected direction of the flow surface at the radially inner edge has a direction component tangential to the rotating shroud and opposite to the rotational direction of the impeller.

In some embodiments, the rotating and inlet shrouds each have at least one extending lip, the lip of the inlet shroud axially overlapping the lip of the rotating shroud to form a static-pressure-reducing labyrinth. An annular space between the lips directs recirculating flow toward the flow control vane.

In some cases the flow surface of the flow control vane defines an arc between the radially outer and the radially inner edges. Preferably, the arc is substantially circular.

In a presently preferred embodiment, the flow surface of the flow control vane is directed, at the radially outer edge, in a substantially radial direction. For some applications, the flow surface of the flow control vane defines, together with a fan radial through the radially inner edge, an exit angle of between about 30 and 60 degrees, preferably about 45 degrees. In some embodiments the inlet is generally circular and axial and has a radius not greater than an inner radius of the rotating shroud. The inlet and rotating shrouds define therebetween, in these embodiments, an exit for recirculating flow re-entering the inlet .

In some configurations, the area of the axial projection of the exit is between about 75 percent and 125 percent of (and preferably about equal to) the area of the axial projection of the inlet annulus .

In some embodiments the radial distance between the radially inner edge and the exit is between about 0.25 and 0.75 times (preferably, about 0.5 times) the radial distance between the radially outer edge and the exit .

In certain arrangements, a recirculation cavity is defined between the rotating and inlet shrouds. The projected area of the flow control vane on a meridional half-plane through its radially outer edge is preferably between about 0.3 and 0.6 times (most preferably, about 0.45 times) the area of the recirculation cavity in the half-plane.

In some embodiments, the recirculation cavity is defined by surfaces of the rotating and inlet shrouds, at least one of which is either partially torus-concave or regionally flat.

In some fans of the invention, the flow control vane is bevelled at a bevel angle, along at least a portion of an edge axially facing the annular space, to present a tapered surface for receiving incoming recirculating flow. The bevel angle is preferably between about 30 degrees and 60 degrees, most preferably about 45 degrees.

In some embodiments, at least a portion of an edge of the flow control vane axially facing the annular space is curved to have a convex surface for receiving incoming recirculating flow. In some instances, the convex surface has a radius at least about equal to the thickness of the flow control vane. In some embodiments, the flow control vane is of molded construction, integral with the inlet shroud.

For some applications, the fan has at least about 25 flow control vanes.

According to another aspect of the invention, a method of circulating air is provided, comprising providing the centrifugal fan described above, and rotating the impeller of the fan with respect to the casing to move air through the fan.

Centrifugal fans with flow control vanes arranged in various ways according to the invention have been demonstrated to provide, in some cases applicable to use in the automotive industry, for example, improvements in output pressure, static efficiency, and/or audible noise, depending on the arrangement . Varying the features of the vanes and recirculation cavity as described herein can improve, for a given application, the output pressure of a fan at a given flow and noise level, or the noise at a given flow and pressure.

Other advantages and features will be apparent from the following description and from the claims.

Brief Description of the Drawing Fig. 1 is a section view of an impeller and inlet shroud of a centrifugal blower of the present invention. Fig. 1A is an enlarged view of area 1A of Fig. 1, illustrating cavity and vane areas.

Fig. 2 is a cross-sectional view of the inlet shroud, taken along line 2-2 in Fig. 1.

Fig. 3 is a cross-sectional view of a flow control vane, taken along line 3-3 in Fig. 2. Fig. 3A is a cross-sectional view of an alternative embodiment of a flow control vane, taken along line 3-3 in Fig. 2.

Fig. 4 shows the flow-redirecting effect of the flow control vanes of the present invention.

Fig. 5 illustrates a second embodiment with additional labyrinth rings.

Fig. 6 is a graph of non-dimensional output pressure v. non-dimensional flow for various fan configurations.

Fig. 7 is a graph of static efficiency v. non- dimensional flow for the fans of Fig. 6.

Description of the Preferred Embodiments Referring to Fig. 1, a centrifugal fan 10 includes a rotating, bladed impeller 12 of about seven inches outer diameter and a stationary housing 14. Although only the inlet portion of housing 14 is shown, the housing, as is typical of conventional centrifugal fan designs, extends around the impeller to generally enclose the hub 22 side of the impeller and to guide or direct the outlet flow from the fan.

Impeller 12 is driven by a shaft 16 to rotate in a predetermined direction. Thirty-seven blades 18 about the circumference of impeller 12 move air from a generally axial inlet flow, indicated by inlet flow arrow A, radially outward, as indicated by outlet flow arrow B. Blades 18 extend, generally in the direction of fan axis 20, from a rotating hub 22 to a rotating shroud 24.

Housing 14 has a stationary inlet shroud 26 to direct the incoming flow to impeller 12. A small gap 28 provides running clearance between rotating shroud 24 and stationary shroud 26. Due to the positive pressure difference between fan outlet and inlet, a portion of the outlet flow recirculates through annular gap 28 to rejoin the inlet flow. Gap 28 is formed between an axially extending lip 30 of stationary shroud 26 and an axially extending lip 32 of rotating shroud 24. The geometry of gap 28 is such that air flow through the gap has very little radial component.

The rotating and stationary shrouds 24 and 26 are shaped to substantially enclose a cavity 34 between them. In this embodiment the facing surfaces 36 and 38 of shrouds 26 and 24, respectively, are partially torus- concave, meaning that they are curved to approximate portions of the inner surfaces of a hollow torus. Inner lip 40 of stationary shroud 26 extends radially inward to radially overlap the inner lip 42 of rotating shroud 24, defining between lips 40 and 42 an exit gap 44 through which recirculation flow re-enters the impeller inlet. The projected area of exit gap 44, on a plane perpendicular to fan axis 20, is preferably within about 25 percent of (most preferably about equal to) the area of the axial projection of gap 28. Sixty-eight stationary flow control vanes 46 extend from shroud 26 into cavity 34 to help redirect the recirculation flow entering the cavity. Vanes 46 are integrally molded with inner surface 36 and lip 30 of stationary shroud 26 and are generally parallel to fan axis 20. The vanes are evenly arranged about the circumference of the shroud and preferably extend far enough to axially overlap a portion of lip 42 of the rotating shroud. Each vane has an exposed inlet edge 48 and outlet edge 50, and the vanes are located such that recirculation flow entering cavity 34 through gap 28 tends to flow from a radially outer portion of inlet edges 48, near gap 28, to outlet edges 50. The velocity of the recirculation flow leaving gap 28 has both an axial component and a swirl component, a tangential velocity in the direction of impeller rotation. Outlet edges 50 are preferably radially located near the middle of cavity 34. In other words, outlet edges 50 are positioned, radially, preferably midway between lips 30 and 40, or at least not closer to either lip 30 or lip 40 than about 25% of the radial distance between lips 30 and 40. This positioning of outlet edges 50 relative to cavity 34 provides sufficient space for the "mixing out" of the flow structure before the flow is reingested by the fan. We have found that it is preferable that vanes 46 are reasonably large as compared to cavity 34. Fig. 1A compares cavity area A_c, with single cross-hatching, to vane area A_v, with double cross-hatching. Both A_c and A are measured as projected on the cross-section as shown (i.e., projected onto a meridional half-plane).

Preferably, vane area A is between about 0.3 and 0.6 times (most preferably 0.45 times) cavity area A_ and the radial projection of the vane (i.e. the length of the projection of edge 48 as shown in Fig. 1) is at least about 3 to 4 times the width of gap 28.

Referring to Fig. 2, each vane 46 is curved in an axial view to form a circular arc having a substantially radial direction at the radially outermost portion of the vane. Due to fan rotation and the action of the fan blading, air flow entering cavity 34 has a tangential velocity component, shown by arrow C (the fan rotation direction is indicated by arrow L) . Recirculation airflow thus tends to impinge the inner surfaces 52 of vanes 46 near their radially outer portions, and travel along inner surfaces 52 toward outlet edges 50, where the flow leaves inner surfaces 52 with an exit flow velocity generally tangential to the inner surface of each vane, as indicated by arrow D. This exit velocity D has a radial component (arrow E) , as well as a tangential component (arrow F) in an opposite direction to inlet tangential velocity (arrow C) . In other words, vanes 46 act to redirect the recirculation flow from having a swirl in the direction of impeller rotation to having a swirl in the direction opposite impeller rotation. The tangential direction of inner flow surface 52 at outlet edge 50 defines an exit angle θ with the radial direction. Exit angle θ is preferably between about 30 and 60 degrees, most preferably about 45 degrees.

Fig. 3 shows the chamfer 54 along the inlet edges of the flow control vanes 46 in their radially outer regions. This chamfer helps to present an aerodynamic wedge to the entering recirculation flow. In some embodiments this effect is achieved by a curved inlet edge 56, as shown in Fig. 3A, in which cases it is preferable that curved edge 56 have a radius r that is at least as large as the thickness t of the vane.

Fig. 4 helps to illustrate the redirection of the recirculation flow by flow control vanes 46 in three dimensional space. Flow approaching the vanes is illustrated by arrow G, and flow leaving the vanes is illustrated by arrow H. Arrow L indicates impeller rotational direction.

Referring to Fig. 5, a second embodiment has an additional set of overlapping lips 58 and 60 and another cavity 62 between rotating shroud 124 and stationary shroud 126, creating a labyrinth to help to reduce the amount of recirculation flow. The vanes 128 of this embodiment are compact versions of vanes 46 in the embodiment of Fig. 1. The arrangement illustrated in Fig. 5 has demonstrated particularly good noise characteristics under certain conditions.

Referring to Figs. 6 and 7, a standard centrifugal-flow impeller was tested with different stationary housings to evaluate the performance characteristics of fans with different configurations of leakage stators . In both figures the performance of the fan without flow control vanes is shown as line S. The performance of the same fan but with radial flow control vanes is shown as line R. The performance of the same fan with pre-swirl inducing flow control vanes, as illustrated in Fig. 4, is shown as line P. In all fans tested, the running clearance between the impeller and the stationary shroud was about 1.4 percent of impeller diameter. For comparison purposes, the performance data of the tested fans are shown herein as a function of non- dimensional flow (J) , which is defined as

J = ^V π nD'

where V = flow rate, n = fan speed (rev/sec) , and

D = fan diameter. Two of the variables presented in Figs. 6 and 7 are non- dimensional static pressure ( -_/S) , defined as

where ΔP_S = rise in static pressure, and p = air density,

Δ P_S-V and 77_s=

Q ( 2πn)

where Q - shaft torque .

As can be seen from Fig. 6 , at a non-dimensional flow (J) of about 0.5 the fan with the pre-swirl leakage stator (line P) produced about a 20 percent increase in pressure (K^ as compared to the same fan with a leakage stator with radial vanes (line R) , and about a 45 percent increase in pressure as compared to the same fan without stator vanes (line S) . The static efficiency of the fan was also increased with the use of the pre-swirl stator, as shown in Fig. 7. At a non-dimensional flow (J) of about 0.5 the efficiency was increased by about 10 percent over both of the other fans tested, while at a higher flow (J) of 0.6, the efficiency was increased by over 20 percent. Noise quality was also improved by the use of preswirl -inducing flow control vanes, as compared to radial vanes. Frequency analysis of the noise produced by each fan indicated smaller energy peaks at blade passing frequency harmonics with preswirl-inducing vanes. Initial testing indicates that not only is the narrow band noise spectrum more uniform (agreeable) , but that the overall broadband noise level is decreased as well. When scaled to the same flow rate, the preswirl fan (line P in Figs. 6 and 7) operated at a lower noise level (64 dBA v. 65 dBA) compared with the fan with radial vanes (line R) while developing the same pressure value.

Although accurate instantaneous flow mapping within the recirculation cavity is difficult at best, the following possible explanation is offered of the mechanics behind these noted improvements. This should only be considered to be one theory of operation, and not as limiting the interpreted scope of the claims.

It is believed that standard (i.e. radial) leakage stator flow control vanes introduce turbulent jet flows whose scale is similar to the spacing between vanes. In the fans of the above tests, this spacing was about 3 millimeters. Blade encounters with this flow structure account for a significant portion of acoustic energy at higher frequencies (e.g. at 2.7, 5.4 and 9.2 KHz in these tests) . The ear interprets this energy as unpleasant high frequency noise. Preswirl -inducing leakage stator flow control vanes, in conjunction with cavity 34, can provide an opportunity for these jet wakes to "mix out" or substantially dissipate before reaching the impeller blades, resulting in lower high frequency noise.

Referring back to Fig. 1, a portion of the noise improvement effect may be due to the fact that the recirculation flow leaving the flow control vanes travels a longer distance before encountering the blades, due to the tangential component of its velocity and the presence of an annular space between exit edge 50 of vanes 46 and exit 44. The exiting flow thus has more distance and time for jet wake structures and turbulence to dissipate before it is reingested by the impeller. The rotational (swirling) nature of this flow may also help to reduce the amplitude of velocity variations.

As to the increase in pressure performance caused by the addition of preswirl -inducing flow control vanes (see, e.g., Fig. 6), this may be explained in part by the increase in reingested recirculation flow velocity relative to the blades, due to the tangential velocity component of the recirculation flow opposite to blade velocity. The number of flow control vanes is preferably greater than the number of impeller blades for general applications, although the actual number of vanes will range anywhere from about 25 to about 100. The embodiment shown in the drawing is typical of fans intended for automotive applications, which currently tend to have fan diameters of about six inches to about eight inches.

Other embodiments are within the scope of the following claims.

Claims

What is claimed is:

1. A centrifugal fan comprising an impeller that rotates in a direction on an axis, the impeller having a rotating hub and rotating shroud and blades extending generally axially from the shroud to the hub, and a casing having an inlet shroud defining an inlet for incoming flow, there being defined an annular space between the rotating shroud and the inlet shroud for running clearance such that recirculating airflow moves through said annular space toward said inlet, the inlet shroud having a flow control vane arranged to redirect the recirculating airflow so as to substantially increase its velocity component in a direction tangential to the impeller shroud and opposite to the rotational direction of the impeller.

2. A centrifugal fan comprising an impeller that rotates in a direction on an axis, the impeller having a rotating hub and rotating shroud and blades extending generally axially from the shroud to the hub, and a casing having an inlet shroud defining an inlet for incoming flow, there being defined an annular space between the rotating shroud and the inlet shroud for running clearance, the inlet shroud having a cambered flow control vane having a flow surface arranged to redirect recirculating flow from said annular space, the flow surface extending from a radially outer edge to a radially inner edge, the projected direction of said flow surface at said radially inner edge having a direction component tangential to said rotating shroud and opposite to the rotational direction of the impeller.

3. The centrifugal fan of claim 1 or 2 wherein the rotating and inlet shrouds each have an extending lip, said lip of the inlet shroud axially overlapping said lip of the rotating shroud to form a static- pressure-reducing labyrinth, said annular space for directing recirculating flow toward said flow control vane being defined between said lips.

4. The centrifugal fan of claim 2 wherein the flow surface of said flow control vane defines an arc between said radially outer and said radially inner edges .

5. The centrifugal fan of claim 4 wherein the flow surface of said flow control vane is directed, at said radially outer edge, in a substantially radial direction.

6. The centrifugal fan of claim 4 wherein said arc is substantially circular.

7. The centrifugal fan of claim 2 wherein the flow surface of said flow control vane defines, together with a fan radial through said radially inner edge, an exit angle of between about 30 and 60 degrees.

8. The centrifugal fan of claim 7 wherein said exit angle is about 45 degrees.

9. The centrifugal fan of claim 1 or 2 wherein said inlet is generally circular and axial and has a radius not greater than an inner radius of said rotating shroud, an exit for recirculating flow re-entering said inlet being defined between said inlet shroud and rotating shroud.

10. The centrifugal fan of claim 9 wherein the area of the axial projection of said exit is between about 75 percent and about 125 percent of the area of the axial projection of said inlet annulus .

11. The centrifugal fan of claim 10 wherein the area of the axial projection of said exit is approximately equal to the axial projection of said inlet annulus .

12. The centrifugal fan of claim 9 wherein the radial distance between said radially inner edge and said exit is between about 0.25 and 0.75 times the radial distance between said radially outer edge and said exit.

13. The centrifugal fan of claim 12 wherein the radial distance between said radially inner edge and said exit is about half the radial distance between said radially outer edge and said exit.

14. The centrifugal fan of claim 1 or 2 wherein a recirculation cavity is defined between said rotating and inlet shrouds, the projected area of said flow control vane on a meridional half-plane through its radially outer edge being between about 0.3 and 0.6 times the area of said recirculation cavity in said half-plane.

15. The centrifugal fan of claim 14 wherein said projected area of said flow control vane is about 0.45 times the area of said recirculation cavity in said half- plane .

16. The centrifugal fan of claim 14 wherein said recirculation cavity is defined by surfaces of said rotating and inlet shrouds, at least one of which is either partially torus-concave or regionally flat.

17. The centrifugal fan of claim 1 or 2 wherein said flow control vane is bevelled at a bevel angle, along at least a portion of an edge axially facing said annular space, to present a tapered surface for receiving incoming recirculating flow.

18. The centrifugal fan of claim 17 wherein said bevel angle is about 45 degrees.

19. The centrifugal fan of claim 1 or 2 wherein at least a portion of an edge of said control vane axially facing said annular space is curved to have a convex surface for receiving incoming recirculating flow.

20. The centrifugal fan of claim 19 wherein said convex surface has a radius at least about equal to the thickness of said control vane.

21. The centrifugal fan of claim 1 or 2 wherein said flow control vane is of molded construction, integral with said inlet shroud.

22. The centrifugal fan of claim 1 or 2 having at least about 25 flow control vanes.

23. A centrifugal fan comprising an impeller that rotates in a direction on an axis, the impeller having a rotating hub and rotating shroud and blades extending generally axially from the shroud to the hub, and a casing having an inlet shroud defining an inlet for incoming flow, there being defined an annular space between the rotating shroud and the inlet shroud for running clearance, said inlet being generally circular and axial and having a radius not greater than an inner radius of said rotating shroud, an exit for recirculating flow re-entering said inlet being defined between said inlet shroud and rotating shroud, the inlet shroud having a cambered flow control vane having a flow surface arranged to redirect recirculating flow from said annular space, the flow surface extending from a radially outer edge to a radially inner edge and defining an arc between said radially outer and said radially inner edges, the flow surface of said flow control vane defining, together with a fan radial through said radially inner edge, an exit angle of between about 30 and 60 degrees, the projected direction of said flow surface at said radially inner edge having a direction component tangential to said rotating shroud and opposite to the rotational direction of the impeller, the rotating and inlet shrouds each having an extending lip, said lip of the inlet shroud axially overlapping said lip of the rotating shroud to form a static-pressure-reducing labyrinth, said annular space for directing recirculating flow toward said flow control vane being defined between said lips.

24. A method of circulating air comprising providing the centrifugal fan of claim 1, and rotating the impeller of the fan with respect to the casing to move air through the fan.