BACKGROUND
This application relates to fan technology for use in forced-air thermal management systems such as for example, in electronics cooling, thermosyphons, outdoor power electronics such as PV solar inverters, EV charging stations, telecom, outdoor electronics enclosures, HVAC systems and for general-purpose fan applications.
Fan-driven forced-air heat transfer systems are the most prevalent thermal management solution for electronics cooling. Recently, markets such as automotive, telecom, PV solar power and ground-based defense electronics are pushing electronics products into increasingly hot and harsh environments which can significantly degrade rotary fan life, since bearing lubricants evaporate exponentially faster as operating temperatures increase. In harsh environments, bearings can also aspirate atmospheric contaminants causing degradation of lubricants, thereby further reducing fan life.
An additional unsolved fan problem pertains to outdoor electronics enclosures used in telecom, renewable energy, electrification, autonomous vehicles, traffic control, et.al. Service providers need to install progressively higher power electronics in these enclosures which requires more fan air flow to dissipate the additional heat. To provide higher flow rates and pressures, rotary fans of a given size must increase their RPM. But since fan noise increases with RPM, rotary fans cannot provide the additional air flow without exceeding industry-imposed noise limits that may exist for these outdoor enclosures. The result is that rotary fan noise can gate end-product performance in many of these strategic tech markets, which can slow adoption of higher power electronics in these key strategic markets.
The primary sources of noise in a rotary fan are fluid flow and the bearings which become noisier with wear. Fluid flow noise becomes louder with increases in blade pass frequency (BPF) and blade tip velocity (BTV). Rotary fans are arguably at the top of their S curve and can now only provide trade-offs between metrics such as flow rate, pressure, life and noise. For example, rotary fan noise can be reduced with changes in fan blade design, but these same blade design changes will reduce air flow rates and pressures. When a product has reached the top of its S curve, it has reached the limit of its architecture's physics. To provide concurrent improvements in multiple fan metrics requires a new architecture governed by a different operational physics.
In order to overcome the previously mentioned limitations of rotary fans and to satisfy the current unmet market needs of extended fan life in hot harsh environments and higher air flow rates and pressures with lower noise levels, the present application is directed to a reciprocating air compression fan (C-Fan) having an architecture that combines fluid dynamic characteristics of centrifugal fans and positive-displacement air-compression machines. When compared to rotary fans, C-Fans provide (1) lower BPF and BTV for a given pressure or flow rate resulting in lower fan noise, thereby enabling higher pressures and flows without exceeding industry-imposed noise limits, (2) higher pressures at a given flow rate due to their reciprocating air compression and (3) extreme life extension due to the absence of rotary bearings.
BRIEF DESCRIPTION OF THE DRAWINGS
The accompanying drawings, which are incorporated in and form a part of the specification, illustrate the various disclosed embodiments and, together with the description, serve to explain the principles of the invention. In the drawings:
FIG. 1 illustrates an exemplary embodiment of a fan blade assembly that includes a rigid fan blade with pivot springs.
FIG. 2 illustrates an embodiment of a fan that includes a single-blade fan with a baffle to minimize air back flow.
FIG. 3 illustrates an embodiment of a two-blade fan with a baffle to minimize air back flow and a compression chamber between the blades that increases the fan's delivered air pressure.
FIG. 4 is a side or end view of the fan of FIG. 3, showing the motor's stator and armatures.
FIG. 5 illustrates the air velocity vectors created by an embodiment of the oscillating fan blade disclosed in the application.
FIG. 6 illustrates an embodiment of a two blade fan, wherein two shells having boundary walls are added to the exterior sides of the two blades to create two exterior compression chambers.
FIG. 7 illustrates an embodiment that is a modification of the two blade fan of FIG. 6, such that the fan includes an internal shell between the blades so that the boundary walls of the shell create two independent internal compression chambers for the two blades.
FIG. 8 shows the fan of FIG. 7 and includes hatching to show the time-varying air exit area during the first half of a cycle.
FIG. 9 shows the fan of FIG. 7 and includes hatching to show the time-varying air exit area during the second half of a cycle.
FIG. 10 illustrates an embodiment of a two blade fan (e.g., the fans of FIG. 6 or FIG. 7) that includes end caps used to close the open end of the compression chambers in order to achieve even higher delivered air pressures.
FIG. 11 illustrates another embodiment of a single blade fan that includes two compression chambers, with one compression chamber located on either side of the fan.
FIG. 12 illustrates another embodiment of a two blade fan with a compression chamber boundary wall.
FIG. 13 illustrates another embodiment of a two blade fan with compression chamber boundary walls.
FIG. 14 illustrates another embodiment of a two blade fan including an option end cap.
DETAILED DESCRIPTION
FIG. 1 provides an example of a fan blade assembly having a stationary spring clamp plate 2, a rigid fan blade 6 and multiple coil springs 4 with one end of each coil spring being attached to stationary clamp plate 2 and the other end of each coil spring attached to fan blade 6. The springs will typically be formed from steel spring wire. In operation, fan blade 6 is free to oscillate back and forth as a rigid panel, like a door on a hinge, in the positive and negative x direction by pivoting on springs 4. The air displaced by the oscillation of blade 6 results in a net air flow in the Y direction.
As is true for rotary fans, the fan disclosed and described herein may include a baffle that maximizes the fan's pressure delivery by minimizing air back flow. The baffle may include a window (e.g., opening or passage) that frames the blade's motion and creates a small clearance between the blade and window thereby creating a pressure seal to the end product.
A fan baffle 8 is shown in FIG. 2 having a window 10 that frames the motion of the blade tip edge 7 with a small clearance between the swept area of the blade tip edge and window 10, thereby preventing contact between the blade tip and baffle 8 while minimizing the back flow area between the moving blade and window. Baffle 8 is rigidly connected to stationary clamp plate 2 by a series of brackets 3. Baffle 8 provides a pressure seal with the product that it is moving air through and creates a boundary between the high-pressure side PH and the low-pressure side PL of the fan.
For the fan blade assemblies of FIG. 1 and FIG. 2, the width W of blade 6 is 19 in, the height HB is 3.9 in and the blade tip to the blade pivot axis HB+HS is 4.5 in. When installed in an ASME fan tester with the baffle 8 sealed to the fan tester flow entrance, the single fan blade of FIG. 2 delivers a zero-pressure flow of 267 CFM through baffle window 10 and a zero-flow pressure PH of 82.3 Pa when operated at 65.6 Hz with a blade tip stroke of 35 mm.
Calculation of the flow rate and pressure for the single blade fan of FIG. 2 can be accurately predicted with the following equations. Equation 1 gives the zero-pressure air flow rate of the FIG. 2 fan, where the factor of 2 accounts for the two blade strokes per cycle, Vs is the blade's swept air volume for a single blade stroke and fo is the fan's operating frequency.
Equation 2 is the peak air velocity exiting the window, based on a sinusoidal exit velocity waveform, where {dot over (Q)} is the average flow rate of equation 1, A is the area of the air flow exit window and the factor of π/2 that converts average velocity into peak velocity occurs when solving the average value integral for a fully rectified sine function, which is a good approximation of the window's exit air velocity waveform.
The max-pressure zero-flow condition is provided by equation 3 where ρ is the density of air.
Equation 3 closely predicts the measured 82.3 Pa zero-flow pressure of the FIG. 2 fan.
FIGS. 3 and 4 show a two-blade fan having blade assemblies 12 and 14 each being identical to the blade assembly of FIGS. 1 and 2. Blade assembly 12 delivers air flow through window 18 of baffle 16 and blade assembly 14 delivers air flow through window 20 of baffle 16. The angle θ between the blades is 13º and the at-rest spacing D between the blade tips is 91.5 mm. Baffle 16 is rigidly connected to fan frame 21 by a series of brackets 17. When installed in an ASME fan tester with the baffle 16 sealed to the fan tester flow entrance, the fan of FIG. 3 delivers a zero-pressure flow of 534 CFM through baffle windows 18 and 20 and a zero-flow gauge pressure PH of 316 Pa when operated at 65.6 Hz with a blade tip stroke of 35 mm for both blade assemblies 22 and 24. The displacements of blades 22 and 24 are 180° out of phase with each other, which means the blades are moving towards each other during one half of the cycle and away from each other during the second half of the cycle. The blade oscillations are driven by a motor as shown in FIG. 4, comprising a stator 23 rigidly mounted to fan frame 21 and armatures 27 mounted to blades 22 and 24. In operation, stator 23 applies time varying electromagnetic forces to armatures 27, thereby causing the blades to oscillate. A motor comprising a stator and armatures as shown in FIG. 4, can be used in any of the embodiments described herein, but is omitted from them for simplicity of illustration.
For the fan of FIG. 3, blades 22 and 24 operate at the same stroke amplitude and frequency as the identical blade 6 of FIG. 2. Two rotary fans operating in parallel will deliver 2× the flow rate of either fan alone but the pressure will be the same as a single fan. The two fan blades of FIG. 3 also deliver 2× the flow of a single fan blade, but unlike rotary fans the fan blades of FIG. 3 deliver 3.84× the measured 82.3 Pa pressure of the FIG. 2 fan. This 3.84× pressure gain of 2 blades vs. a single blade is explained as follows.
The embodiments of the fans disclosed herein operate at their mass-spring mechanical resonance, which results in a sinusoidal oscillation of the fan blade. FIG. 5 illustrates the air velocity vectors of a single blade fan. This velocity vector representation is commonly used in the theory of centrifugal fans. The velocity triangle represents the instantaneous air velocity vectors during the half cycle when blade 6 is moving from left to right. Vector VT is the air velocity tangent to the arc of the blade tip's displacement, vector VP is the air velocity that leaves the tip of the blade in a direction parallel to the blade and vector VR is the resultant air velocity direction and magnitude which is the vector sum of VT and VP. Velocity vector magnitudes vary sinusoidally with the blade's sinusoidal velocity. Vector VT will always point in the direction of the blade's motion. The vector triangle of FIG. 5 is for illustrative purposes only and the actual relative magnitudes of VT and VP will be a function of a given fan's geometric design and operating parameters.
The 2-blade fan design of FIGS. 3 and 4 enable the added air pressurization of a reciprocating-volume, or positive displacement, machine to the kinetically-driven pressure of the FIG. 2 single-blade fan. When blades 22 and 24 are moving towards each other the rapidly reducing mechanical volume 25 bounded by the blades 22 and 24 creates an additional time-varying cyclic pressure in the air within volume 25. The peak pressure within the 2-blade bounded volume 25 can reach much higher values than the pressures created by the single-blade case, where the air in contact with either side of the blade is unbounded. In FIG. 3, the instantaneous velocities measured at the window exits of baffle 16, are the superposition of (1) the blade's kinetically-driven velocity as represented in FIGS. 5 and (2) the additional reciprocating-volume driven velocity of the compressed air within volume 25. The distance D between the blades will affect the reciprocating pressure magnitude developed within volume 25. Assuming the baffle windows remain centered on the blades, the air pressure amplitude within volume 25 will decrease as D increases. As D is increased for the 2-blade fan, the total pressure delivered will approach the pressure delivered buy a single blade fan having the same blade size, frequency and stroke. In other words, as D is increased a value of D will be reached where the two fan blades will behave as though the air around the blades is unbounded as in the case of the single blade fan of FIG. 2. Further, the angle θ between blades 22 and 24, as shown in FIG. 4, is an important design parameter for fan designs when the blades 22 and 24 share the same compression volume 25, whereby θ can be varied to optimize the peak pressure created within volume 25 for given values of D and blade stroke.
Further increases in pressure can be provided if a reciprocating-volume exists for both halves of the cycle. This can be accomplished by adding a shell on the outside of the blades as shown in FIG. 6, where an outer shell 26 is provided for blade 22 and an outer shell 28 is provided for blade 24. Shells 26 and 28 will typically have a width being equal to or greater than the width W of the blades and can be formed from sheet metal in the V-shape to improve structural integrity of the part. The volume within the shells plays no role in the functioning of the fan. The surfaces 27 and 29 of respective outer shells 26 and 28 form the exterior boundary walls of respective bounded air volumes 30 and 32 on the outside of respective blades 22 and 24, which creates air compression within bounded volumes 30 and 32 during the half cycle when the blades are moving away from each other. The fan design of FIG. 6 is identical to that of the FIG. 3 fan except for the addition of the external compression volumes. When operating at the same condition as the FIG. 3 fan (65.6 Hz, 35 mm blade tip stroke), the FIG. 6 fan will deliver the same open flow rate of 534 CFM but a higher zero flow pressure of 475 Pa, which is a 5.8× gain in pressure compared to a single blade fan without compression volumes. Pressure gain values much higher than 5.8× can be achieved and are a just function of the geometrical design parameters of a given fan.
In FIG. 7, an inner shell 34 is added between the blades of the fan of FIG. 6. Inner shell 34 separates the inner compression volume 25 of FIG. 6 into two separate and independent inner compression volumes 36 and 38 for respective blades 22 and 24. Inner shell 34 can be formed from sheet metal as shown in FIG. 7 to provide structural integrity but the volume within inner shell 34 plays no role in the functioning of the fan. The surfaces 33 and 35 of inner shell 34 form the interior boundary walls of respective bounded air volumes 36 and 38 on the inside of respective blades 22 and 24. Separate inner compression volumes essentially creates two fluid dynamically independent fans that are no longer dependent on the blade-to-blade distance D shown in FIG. 4. Compression shells can be used on one side or both sides of a single blade flan to achieve higher pressures independently from having a second blade. For example, FIG. 11 shows a fan having a fan blade assembly 50, baffle 56, baffle window 58, compression shells 52 and 54 rigidly connected to baffle 56 and having a first boundary wall 53 and a second boundary wall 55 of the respective bounded volumes 60 and 62 and brackets 64 that rigidly connects the blade assembly to the baffle.
A fan blade of given dimensions and operating at the same stroke and operating frequency will deliver the same zero-pressure air flow rate (aka open flow) either with or without a reciprocating compression volume, since open flow is determined solely by the volumetric displacement rate of the fan blades. However, when a reciprocating compression volume is added, the same fan blade, stoke and operating frequency will deliver higher pressures at a given air flow rate. The explanation for this pressure behavior is provided as follows.
Referring to FIG. 8, during the half cycle when blades 22 and 24 are moving from positions A to C, the air displaced by blades 22 and 24 will be pressurized within respective air volumes 30 and 32 and will exit through the respective hatched window areas 40 of window 18 and 42 of window 20. Since blades 22 and 24 are oscillating sinusoidally, their driven instantaneous air velocities are being delivered sinusoidally. As blades 22 and 24 begin their travel from positions A, their driven instantaneous air velocities increase from zero at positions A to their peak value at positions B and then begin decreasing to zero at positions C.
As blades 22 and 24 move from positions A to C the resulting positive displacement compression in respective air volumes 30 and 32 increases the pressures in air volumes 30 and 32 to levels that are higher than those that could occur without pressure volume 30 and 32, as illustrated for example by the fan of FIG. 3. These higher pressures create higher air velocities through the decreasing window areas 40 and 42, resulting in higher average velocities and therefor higher delivered fan air pressures during this half cycle.
As illustrated in FIG. 9, during the second half of the cycle when blades 22 and 24 are moving from positions C to positions A, the process repeats with air volumes 36 and 38 being pressurized and air flow exiting through hatched areas 44 and 46. Just like the A-to-C cycle, the C-to-A cycle results in higher average air velocities and higher resulting delivered fan air pressures during this half cycle. The fan of FIG. 6, works according to the same principles as the fan of FIG. 9 with the only exception being that the C-to-A cycle employs a shared compression volume 25 for the fan of FIG. 6 instead of the separate compression volumes 36 and 38 of the FIG. 8 and FIG. 9 fans.
Adding compression volumes does not change the open flow rate for a given blade stroke and frequency operating condition, but it does result in higher average velocities and higher resulting pressures for the same operating condition with compression volumes removed. This added air compression will occur with the fans shown in FIGS. 3, 4, 6, 7, 8 and 9 even though the compression volumes are open to the ambient air at the far ends of their blade's width W. These open ends of the compression chambers can be closed as shown in FIG. 10 with a compression chamber end cap 48, which can further improve the pressure gain. When using a compression chamber end cap, care must be taken in the fan design to leave adequate air inflow area into the compression chambers.
Compression chambers are created when a boundary wall is added on either side of the fan blade and the design of the boundary wall will be dependent on the requirements of a given fan application. For example, FIG. 12 shows the outer compression chamber boundary walls 68 and 70 and an inner compression chamber boundary wall 80 being formed of single sheets rather than the shells shown in other embodiments.
Baffles can increase a fan's delivered air pressure by minimizing reverse air flow, as described herein, but the design of the baffle can vary based on factors such as the design of the compression chambers. As illustrated in FIG. 13, stationary boundary walls 82 and 84 of respective compression chambers 88 and 90 can create the same time-varying exit area as a baffle and the same reduction is reverse air flow area. It can also be appreciated that the single sheet boundary wall 86 could be removed in the fan of FIG. 13, without increasing reverse air flow due to the close approach of the fan blade tips. A baffle designed for the fan of FIG. 13 would provide a pressure seal to boundary walls 82 and 84 and to the product receiving the fan's air flow. Optional end caps, as shown in FIG. 10, could also be added to the FIG. 14 fan. Whenever a fan employs a compression chamber and a baffle, such as described in the various embodiments disclosed herein, air is free to flow into the compression chamber from the low-pressure side of the fan and flow out of the compression chamber through the window in the baffle and into the high-pressure side of the fan.
The various structures of fans disclosed herein may also be used as a form of propulsion in air or in liquids. For example, the structures described herein could be modified to function as propulsion motors or engines for various air, land and water based vehicles.
The foregoing examples have been provided merely for the purpose of explanation and are in no way to be construed as limiting of implementations of the present invention.
Patent Application
20240426307
