Patent Application
20100329871
The present invention relates to fans and fan assemblies suitable for automotive applications.
Modern vehicles, such as medium- and heavy-duty diesel trucks, can have relatively high cooling demands. For instance, diesel engine emissions requirements mandated by European and North American regulations have placed greatly increased demands upon engine cooling systems. Not only is more airflow required to provide adequate cooling and increased pressure required to overcome the restriction of radiators and other heat exchangers, but vehicle designs dictate and limit the size of cooling system components. Such limitations are of particular concern when low hood lines are desired with truck and construction equipment for better driver visibility. Without being able to increase an exposed surface area of radiators and other heat exchangers, they are often made thicker. Thicker (i.e., deeper) radiators and other heat exchangers reduce engine compartment space available for other cooling system components, such as fans and fan clutches.
Automotive applications have traditionally employed axial flow fans to provide cooling flows. Axial flow fans generally move air in a direction parallel to an axis of rotation of the fan. However, the combination of increased flow requirements and thicker heat exchangers radically increases the restriction of cooling systems, to the point where conventional axial flow fans are no longer capable of providing an adequate flow of air. Even with fan systems that can be enlarged, the relatively low efficiency of conventional axial flow fans cause excessive power draws (e.g., greater than or equal to about 15% of engine power) that reduce useable power from the engine. Moreover, axial flow fans may not operate as quietly as desired for automotive applications, which can be a concern for meeting noise regulations.
It is well-known that mixed flow fans (also known as hybrid flow fans) and radial flow fans (also known as centrifugal fans) have greater efficiencies and flow-pressure characteristics than axial flow fans, but mixed flow and radial flow fans are difficult to package in most vehicle engine compartments. Radial flow fans typically require large scroll housings for best efficiency, and if used without such housings have radial discharge velocities that are not conducive to movement around vehicle engines. Although mixed flow fans do not have those problems of radial flow fans, they are typically thicker (i.e., deeper) in the axial direction than can be used in under-hood applications. Furthermore, mixed flow fans are deceptively complicated devices. While the general idea of a mixed flow fan appears simple, the tremendous amount of experimentation and design required to tailor them to meet the requirements of particular applications has meant that they are rarely used in practice.
A fan assembly for directing fluid flow in a hybrid radial and axial direction includes a backplate having an inner diameter portion and a substantially frusto-conical outer diameter portion positioned about a center axis, a plurality of blades extending from the backplate, and an annular fan shroud positioned adjacent to the plurality of blades and configured for co-rotation therewith. The backplate, the plurality of fan blades and the fan shroud form a fan subassembly, and an overall depth of the fan subassembly is approximately 20-35% of an overall fan subassembly diameter.
While the above-identified drawing figures set forth several embodiments of the invention, other embodiments are also contemplated, as noted in the discussion. In all cases, this disclosure presents the invention by way of representation and not limitation. It should be understood that numerous other modifications and embodiments can be devised by those skilled in the art, which fall within the scope and spirit of the principles of the invention. The figures may not be drawn to scale. Like reference numbers have been used throughout the figures to denote like parts.
The present invention claims priority to U.S. Provisional Patent Application No. 61/066,692 entitled “High Efficiency Hybrid Flow Fan,” filed Feb. 22, 2008, which is hereby incorporated by reference in its entirety.
In general, the present invention provides a quasi-mixed (or hybrid) flow fan (generally referred to herein simply as a hybrid flow fan), enabling the generation of fluid flow in a hybrid radial and axial direction (i.e., somewhere in between 0 and 90° with respect to the axial direction) in response to rotational input. In one embodiment, the fan has an overall depth (i.e. thickness or width) of approximately 20-35% of an overall fan diameter. The fan of the present invention can be used in engine cooling systems, preferably when operating in a range of fan throttling coefficients from approximately 0.04 to 0.08, where throttling coefficient is defined as a ratio of velocity pressure to total pressure, with the velocity pressure calculation based on a superficial velocity equal to airflow divided by an axial projected area of the fan.
The fan of the present invention provides numerous advantages and benefits. For example, the fan provides a relatively high airflow and relatively high pressure fan for engine cooling. However, configuration of the fan is generally subject to several constraints for use with automotive and other engine cooling applications. The fan should preferably be mounted on the front of an engine in the same manner as existing axial flow fans (e.g., belt-driven or crankshaft mounted). Further, the fan should allow use of a viscous fan clutch (also called a viscous fan drive), a device that allows speed control of the fan and helps isolate the fan from crankshaft torsional vibration. An overall diameter of the fan should preferably be comparable to existing axial flow fans. A thickness (i.e., axial depth) of the fan should ideally be comparable to existing axial flow fans, or as thin (i.e., axially narrow) as possible because additional engine compartment space is often difficult or impossible to allocate. An inlet diameter of the fan should preferably be as large as possible to prevent high high-velocity airflows in the center of radiators or other heat exchangers that can result in detrimental airflow stratification through radiator and heat exchanger cores. Airflow discharge from the fan should preferably have an axial component to help guide the air around sides of and past the engine. Static efficiency of the fan should be as high as possible, and preferably greater than 50%, to maximize the engine power available for useful work. Noise produced by the fan should be as low as possible, and preferably no louder than that of existing axial-flow fans operating with lesser aerodynamic performance. Also, an interface (i.e., shrouding) between an inlet to the fan and the radiator or other heat exchangers should accommodate relative motion between the two caused by engine rocking and frame twisting, yet be made of structures achievable by ordinary assembly-line procedures.
Several of the constraints discussed above appear mutually exclusive. The inlet diameter of the fan is one such example. Generally, in a radial flow (or centrifugal) fan, greater pressure production is achieved by decreasing a ratio of blade inside diameter to blade outside diameter, thus making fan blades longer in a radial direction. Doing so, however, decreases an axial inlet area of the fan, increasing inlet velocity. Because spacing between a vehicle radiator (or other heat exchanger) and fan is typically short, such high velocity fluid flow directly in front of the fan would likely create undesirable “dead zones” in corners of the radiator (or other heat exchanger), thereby decreasing overall heat exchange efficiency. Similarly, high airflow in a radial flow (or centrifugal) fan is typically achieved by increasing the fan's axial depth, an option not available for under-hood engine cooling applications. It was necessary, therefore, in designing the fan of the present invention to create a fan with design parameters that produced a suitably efficient fan under a host of constraints. In general, the fan of the present invention tends to exhibit relatively high airflow and static efficiency characteristics while still satisfying the constraints discussed above.
Those of ordinary skill in the art will appreciate that in one embodiment the fan apparatus 20 is attached to a suitable clutch (not shown), such as a viscous clutch of the type disclosed in PCT Published Application No. WO 2007/016497 A1, and in turn operatively connected to an engine (not shown). The clutch is typically removably secured to the backplate 22 of the fan apparatus 20 with bolts or other suitable attachment means. The engine and clutch can selectively rotate the fan apparatus 20 at a desired speed, with the fan apparatus 20 moving air to help cool the engine. In a typical application, the fan apparatus 20 is positioned between a radiator and/or other heat exchangers (not shown) and the engine, with fan operation both directing cooling air to the engine and moving air through the radiator (and/or other heat exchangers) to further provide cooling.
The backplate 22 includes a substantially planar inner diameter (ID) portion 34 (also called a hub) and a frusto-conical outer diameter (OD) portion 36. The ID portion 34 is arranged generally perpendicular to the centerline CL of the fan apparatus 20. A metallic disk 38 (e.g., made of steel, aluminum, etc.) is optionally incorporated into the ID portion 34 at the centerline CL to provide a relatively rigid structure for attachment of the fan apparatus 20 to a clutch or other rotational input source (not shown). One or more openings are optionally provided in the metallic disk 38 in the ID portion 34 at or near the centerline CL to facilitate attachment to the clutch or other rotational input source. The ID portion 34 is sufficiently large to accommodate attachment to a clutch. Prior art mixed flow fans tend to have an ID portion that is too small for mounting to a conventional automotive fan clutch. The OD portion 36 is positioned directly adjacent to and radially outward from the ID portion 34. The OD portion 36 is arranged at an angle θ1 with respect to the centerline CL. Generally, a discharge angle of the airflow 33 exiting the fan apparatus 20 is equal to the angle θ1. In the illustrated embodiment, the OD portion 36 extends to a perimeter (i.e., circumference) of the fan assembly 20. The backplate 22 has a radius R1, which defines a corresponding overall diameter øD1. For common applications, values of the diameter øD1 range from about 450 mm to about 750 mm, though it will be appreciated that the diameter øD1 can have essentially any value greater than zero as desired for particular applications.
In the illustrated embodiment, a groove 39 is formed in the rear side of the backplate 22 corresponding to and aligned with each one of the blades 24. The grooves 39 help reduce thickness of the backplate 22 and an overall mass of the fan apparatus 20. The grooves 39 are optional, and generally are only present when the backplate 22 and the blades 24 are integrally molded during fabrication. When the backplate 22 is injection molded, the grooves 39 also help avoid sink marks, which are molding defects that occur due to volume shrinkage during cooling. Fabrication of the fan apparatus 20 is discussed further below.
An annular rib 40 extends generally axially from the backplate 22 at a rear side of the backplate 22 opposite the blades 24 (see
Turning again to
The blades 24 extend from the OD portion 36 of the backplate 22 to the fan shroud 26. In the illustrated embodiment, a total of sixteen blades 24 are provided, though the number of blades 24 can vary in alternative embodiments (e.g., a total of eighteen blades 24, etc.). Each blade 24 defines a leading edge 44, which is oriented at an angle θ3 relative to the OD portion 36 of the backplate 22, and a trailing edge 46, which is arranged substantially parallel to the centerline CL in the illustrated embodiment. Those skilled in the art will appreciate that opposite pressure and suction sides of the blades 24 extend between the leading and trailing edges 44 and 46. In the illustrated embodiment the leading edges 44 of the blades 24 are not attached to the fan shroud 26. The leading edges 44 of the blades 24 collectively define a radius R3 about the centerline CL, which corresponds to a blade inner diameter øD3. Because the blades 24 extend along the frusto-conical OD portion 36 of the backplate 22, the radial locations of the leading edges 44 of the blades 24 affect the center of mass of the fan apparatus 22 in the axial direction. It is generally desirable to locate the center of mass at an axially middle location to better balance the fan apparatus 20 during operation, particularly with respect to bearings of a clutch to which the fan apparatus 20 can be mounted. In some embodiments, the ID portion 34 is substantially aligned with the center of mass of the fan apparatus 20 (e.g., within approximately +/−2% of the overall diameter øD1 relative to the center of mass in the axial direction). Furthermore, each blade defines an inlet angle β1 and an exit angle βE (see
The blades 24 in the embodiment of the fan apparatus 20 shown in
As shown in
The fan apparatus 20 defines a projected width PWf (i.e., an overall depth or thickness) in the axial direction. In the illustrated embodiment, the projected width PWf is defined between the axially forward extent of the fan shroud 26 and an axially rear extent of the OD portion 36 of the backplate 22. In one embodiment, the overall diameter øD1 of the fan apparatus 20 is approximately 550 mm and the projected with PWf of the fan apparatus 20 is approximately 165 mm. While the fan apparatus 20 is generally thicker (i.e., deeper in the axial direction) than a conventional axial flow fan, the fan apparatus 20 can have a thickness of only about 180-200% relative to the thickness of a conventional axial flow fan compared to about 250% for prior art mixed flow fans and about 300% for prior art radial flow fans.
The inlet shroud 32 is an annular member positioned adjacent to the fan apparatus 20, and includes an ID portion 50 that is at least partially curved in a toroidal configuration. The inlet shroud 32 defines an upstream opening that is larger than a downstream opening. Typically, the inlet shroud 32 is rotationally fixed, and in under-hood applications can be secured to an engine, a radiator or other heat exchanger, a vehicle frame, etc. The inlet shroud defines a radius R4 at a radially inward extent of the ID portion 50, with the radius R4 corresponding to a diameter øD4. In the illustrated embodiment, at least part of the ID portion 50 of the inlet shroud 32 is positioned within an upstream portion of the fan shroud 26, and extends rearward of the axially forward extent of the fan shroud 26. In other words, an axial overlap is formed between the fan shroud 26 and the inlet shroud 32. A generally radial gap is present between the fan shroud 26 and the inlet shroud 32, which, in under-hood applications, allows for relative movement between those components due to engine rocking, frame twisting, vibration or other movements. During operation, fluid flow in the direction of the arrow 33 passes through a central opening of the inlet shroud 32 to the fan apparatus 20. The inlet shroud 32 can help guide airflow to the fan apparatus 20 from a radiator or other heat exchanger. Also, some additional fluid flow may reach the fan apparatus 20 through the generally radial gap between the fan shroud 26 and the inlet shroud 32.
The configuration of the fan apparatus 20 according to the present invention can vary as desired for particular applications. Table 1 provides three possible ranges for parameters of the fan apparatus 20. The values given in Table 1 are all approximate. It should also be noted that the values in Table 1 are provided merely by way of example and not limitation. Moreover, Table 1 should be interpreted to allow independent selection of individual parameters. For instance, one parameter can be selected from the “first range” column while another parameter can be selected from the “second range” column, and so forth.
The fan assembly 30, including the fan apparatus 20, can be manufactured in a variety of ways. Typically components of the fan assembly 30 are made of a polymer or other injection-moldable material, though fiberglass, metals and other suitable materials can alternatively be used. In one embodiment, injection molding is utilized, in which a polymer material, such as nylon, forms essentially all of the components of the fan assembly 30, except for the metallic disk 38, which can be made of steel. The blades 24 and the backplate 22 are usually integrally formed as a single subassembly. If the blades 24 and backplate 22 are injection molded, the metallic disk 38 can be overmolded with the polymer material to integrally form the blades 24 and the backplate 22. The fan shroud 26 and the inlet shroud 32 are generally each separately formed by injection molding or other suitable techniques. The fan shroud 26 is then attached to the blades 24 of the subassembly, using a welding process, mechanical fasteners or other suitable techniques. A welding or welding-like process, such as ultrasonic welding or high frequency electromagnetic welding and bonding, is preferred. A configuration with welded joints between the blades 24 and the fan shroud 26 produces relatively low stresses on the weld joints between the blades 24 and the fan shroud 26, while simplifying the process of injection molding the individual parts that are later welded together. The inlet shroud 32 is separately attached to a mounting structure, and the fan apparatus 20 is positioned adjacent to the inlet shroud 32 in a desired installation location.
In other embodiments, the backplate 22, the blades 24 and the fan shroud 26 of the fan apparatus 20 are integrally molded as a single piece. While a single-piece construction offers strength benefits, it tends to require complex and expensive dies to achieve. Alternatively, the fan shroud 26 and the blades 24 are integrally molded and attached to a separately molded backplate 22.
As previously mentioned, a fan apparatus according to the present invention can have its blades arranged in a number of different configurations in alternative embodiments, such as backward curved, backward inclined, radial (or quasi-radial) tip, forward curved, and radial blade configurations. Those terms are derived from radial flow fan design. Different blade configurations will have different operational effects, which are generally interrelated to other fan apparatus parameters. The optimal blade configuration will vary for different applications depending on the desired performance characteristics and constraints on the design of the fan apparatus.
In view of the foregoing description, those skilled in the art will recognize that a fan assembly according to the present invention provides numerous advantages and benefits. For example, a fan according to the present invention provides relatively high pressure and airflow but is relatively thin and generally exhibits a different aspect ratio than what a designer would otherwise produce with the luxury of substantial axial depth space available. Moreover, the fan of the present invention exhibits relatively good operating static efficiency characteristics. The fan of the present invention can also meet desired performance characteristics for under-hood automotive cooling applications while simultaneously satisfying the many design limitations associated with under-hood applications.
In addition, a fan according to the present invention provides relatively good noise characteristics, including both noise intensity and noise quality characteristics. The fairest comparison of noise between two fan types is when both are operating at the same aerodynamic point (i.e. same flow and pressure). Comparing a 680 mm diameter fan of the present invention running 1900 RPM to a prior art 750 mm diameter axial flow fan running at 1970 RPM, the fan of the present invention was 4 dBA quieter. The fan of the present invention is quieter for two major reasons. First, the fan of the present invention can develop a desired level of static pressure at a slower rotational speed compared to an axial flow fan, and fan noise is very strongly dependent upon peripheral speed (i.e., tip speed). Second, flow of air through passages of the fan of the present invention is much smoother and much less turbulent than the flow of air through an axial flow fan at the high pressures at which the fan of the present invention is desired to operate. Typically, flow through an axial flow fan under the conditions described above is known as stalled flow, which is highly turbulent and unstable, and is associated with a roaring noise.
Additional advantages and benefits not specifically mentioned are also provided.
Prototype fan assemblies according to the present invention were developed and tested, and computer simulations were run to further explore fan assembly designs according to the present invention. Prototype testing has shown that a fan according to the present invention can achieve about 35% higher airflow, 15 percentage-points greater static efficiency and exhibit quieter operating characteristics than state-of-the art axial flow fans, while still being suitable for installation in under-hood automotive cooling applications and exhibiting acceptable power requirements.
A design of experiments (DOE) protocol was employed to run simulations of a number of permutations of a number of judiciously selected fan design variables. The DOE allows for optimization while conducting tests on only a limited number of possible permutations. Computational fluid dynamics (CFD) software (e.g., FLUENT® flow modeling software available from ANSYS, Inc., Santa Clara, Calif.) was utilized to generate simulation test data according to each DOE. Multiple DOE studies were conducted. The largest DOE conducted involved five factors with three possible levels each, for a total of 243 (or 35) possible combinations, of which 27 variations were simulated in accordance with the selections of factors and levels listed in Table 2.
Results of the DOE were gathered for airflow rate (in kg/s), static pressure (in Pa) and static efficiency (in %).
The results for pressure vs. pressure vs. airflow data points (solid diamonds) were specified to fall upon a quadratic curve that approximates a typical engine cooling restriction curve. The DOE results show that the corresponding static efficiency vs. airflow data points (hollow squares) collectively define a boundary curve 400. Based on the 27 DOE results, data points were interpolated for three optimized designs of the fan apparatus 20. For a design #1, performance was optimized for both best airflow and best static efficiency, illustrated in
Although the present invention has been described with reference to preferred embodiments, workers skilled in the art will recognize that changes may be made in form and detail without departing from the spirit and scope of the invention.
| Filing Document | Filing Date | Country | Kind | 371c Date |
|---|---|---|---|---|
| PCT/US09/01047 | 2/19/2009 | WO | 00 | 8/16/2010 |
| Number | Date | Country | |
|---|---|---|---|
| 61066692 | Feb 2008 | US |
