Electrical generator or motor structure – Dynamoelectric – Reciprocating
Reexamination Certificate
2000-03-08
2001-10-23
Ramirez, Nestor (Department: 2834)
Electrical generator or motor structure
Dynamoelectric
Reciprocating
C310S017000, C324S076490, C359S199200, C062S006000, C062S467000
Reexamination Certificate
active
06307287
ABSTRACT:
FIELD OF THE INVENTION
In general this invention relates to the use of electrodynamic motors to create high-power, high-efficiency loudspeakers for generating high-amplitude waves in devices such as thermoacoustic refrigerators or acoustic compressors.
BACKGROUND OF THE INVENTION
Over the past fifteen years, a new class of refrigerators and heat engines has been developed. These devices utilize intrinsically irreversible thermal conduction or acoustical geometry to provide the proper phasing between pressure and volumetric velocity. This phasing produces useful quantities of heat pumping or cooling or generates mechanical work. These new engines are called thermoacoustic engines. Earlier engines required mechanical means, such as pistons, linkages, displacers, cams, valves and other mechanisms to realize useful cooling or produce mechanical work using more traditional reversible heat engine cycles, e.g., Stirling Cycle or Rankine Cycle. The heat pumping power of these new acoustic refrigerators depends upon the square of the acoustic amplitude—a doubling of acoustic pressure amplitude corresponds to four times the useful heat-pumping power. It is therefore important to be able to produce very high amplitude sound waves for use in such refrigeration devices.
Over the last decade, a new fluid pump that employs high-amplitude acoustic standing waves has also been developed, U.S. Pat. No. 5,319, 938 and U.S. Pat. No. 5,515,684. These sound waves actuate reed valves that rectify oscillatory pressure to produce the mean flow and to permit continuous, unidirectional pumping of gases by the high-amplitude sound waves. A very high-power sound source is necessary to make this “sonic compressor” effective, and a highly efficient conversion from electrical to acoustical energy is required to make it economical.
Nearly all of the electrically driven thermoacoustic refrigerators patented and/or produced to date have used a moving-coil, electrodynamic loudspeaker to generate the required high-amplitude sound waves. These moving coil loudspeakers had several attractive features associated with their relatively low moving (dynamic) mass. The low mass meant that a fairly flexible suspension could be used to provide a high resonance frequency, usually in the range of several hundred Hertz. This lower moving mass also permitted the operating frequency of the thermoacoustic refrigerator to be adjusted over a modest range of frequencies to allow the system to be tuned over a small bandwidth without substantially degrading efficiency. Unfortunately, the moving-coil loudspeaker efficiency and power handling capacity is limited by the mass of conductor (typically copper, aluminum, or copper-cladded aluminum) in the coil.
The development of high flux density rare-earth magnetic materials (e.g., NdFeB) and the recent invention of a high-efficiency linear alternator which uses such magnets for mechanical-to-electrical power conversion, has made it practical to consider a moving-magnet electrodynamic system as a possible high-amplitude, high-efficiency electrodynamic sound source. The Yarr/Corey design, U.S. Pat. No. 5,389,844, uses several coils wound around a multi-pole laminated magnetic stator to increase the available mass of conductor by several orders-of-magnitude over the moving-coil loudspeaker without directly affecting the length of the magnetic gap or adding to the moving mass which, in the new moving-magnet configuration, is controlled by the mass of the moving rare-earth magnets.
An additional advantage to making the coil part of the magnetic stator is that the electrical leads which bring current to the coil do not have to flex to accommodate motion ofthe moving-magnet part of the motor. The flexure of the input current leads in a moving-coil electrodynamic motor is a fairly common cause of motor failure which the moving-magnet design avoids, thereby increasing its reliability over the moving-coil design.
To appreciate the utility of increasing the operating (mechanical resonance) frequency of such a moving-magnet driver by increasing its resonance frequency by the methods taught herein, one need only to examine the expression for the time-averaged acoustic power, Π
ac
, supplied by a moving piston of area, A
piston
:
∏
ac
=
F
1
v
1
2
=
p
1
A
piston
v
1
2
=
p
1
U
1
2
=
p
1
A
poston
ω
d
1
2
(
1
)
The variables sub-scripted with a “1” are the peak values of quantities which are assumed to have a sinusoidal time variation with a frequency f=&ohgr;/2&pgr;. Observing that convention, the piston stroke (peak-to-peak displacement amplitude) is given by 2d
1
. Further calculation is simplified by the introduction of the piston's peak volumetric velocity, U
1
[m
3
/sec].
The above expression, Eq. 1, includes the simplifying assumptions that the piston speed, V
1
, and the net force on the face of the piston, F
1
, are in-phase. Although there are many circumstances of practical interest for which this assumption is not valid, the assumption is true for a piston that is driving an acoustic load which is oscillating at one of its acoustic resonance frequencies. A resonant acoustic load is commonly found in thermoacoustic refrigerators and sonic compressors.
The acoustic pressure amplitude, p
1
, can be related to the volumetric velocity of the piston, U
1
, by the introduction of an acoustic impedance, Z
ac
[Newton-sec/m
5
], which is given by the ratio of the pressure to the volumetric velocity at the piston location.
Z
ac
≡
p
1
U
1
(
2
)
The acoustic impedance is a function only of the acoustic load presented to the piston and not a function of the drive mechanism.
Substitution of Eq. 2 into Eq. 1 demonstrates that the acoustic power delivered by the piston to the resonator characterized by Z
ac
, is dependent upon the square of the product of the piston displacement, d
1
, and the radian frequency, &ohgr; [rad/sec], of the piston's oscillation.
∏
ac
=
p
1
U
1
2
=
Z
ac
2
U
1
2
=
Z
ac
2
(
d
1
ω
)
2
A
piston
2
(
3
)
It is clear that the full exploitation of moving-magnet electrodynamic motors for the generation of high-amplitude sound fields, with high electroacoustic efficiency, requires that the motors operate at the highest possible frequency, &ohgr;, while maintain the ability to utilize their maximum allowable stroke, 2d
1
. Since the resonance frequency of such a motor, having a moving mass, m
0
[kg], is determined by the total suspension stiffniess, k [Newton/m],
ω
=
k
m
0
(
4
)
it is desirable to increase the total suspension stiffness k without undue restriction on the motor stroke and with an effectively infinite fatigue lifetime for the spring.
The most common approach for providing auxiliary stiffness to increase the resonance frequency of a moving-magnet motor, commonly employed by the Stirling-Cycle engine community, is the use of a “gas spring.” The advantage to the gas spring is that there are no material limitations, such as fatigue fracture, when gas pressure provides the additional restoring force (stiffness). One disadvantage to the gas spring is that there is irreversible thermal dissipation due to the adiabatic heating and cooling of the gas which accompanies the gas spring's compressions and expansions. This dissipation mechanism increases the motor's effective mechanical resistance, R
m
, in addition to increasing its net suspension stiffness, k. Another disadvantage is that the gas spring stiffness is also dependent upon mean pressure, p
m
, in the “back volume” which creates the gas spring, so the driver mechanical resonance will change if the mean operating pressure is changed.
An even more serious limitation to the gas-spring approach is that pressure behind the piston and in front of it are 90° out-of-phase if the driver is located at the pressure anti-node of the standing wave, or nearly 180° out-of-phase for the displaced driver location (located closer to the
Garrett Steven L.
Keolian Robert M.
Smith Robert W.
Gifford Krass Groh Sprinkle Anderson & Citkowski PC
Jones Judson H.
Ramirez Nestor
The Penn State Research Foundation
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