Vibratory acoustic compressor

An acoustic compressor is disclosed for generating a high pressure acoustic waveform within a chamber containing a working medium. The compressor comprises an actuator body solidly mounted to the chamber with a vibrating armature spring-mounted within. The chamber and actuator body are resiliently mounted so as to be vibratable. Owing to reaction forces, vibration of the armature causes vibration of the chamber which generates an acoustic waveform in the medium. By matching the resonant frequency of the armature to the acoustic resonant frequency of the chamber and driving the armature at that frequency, extremely high pressures within the chamber may be obtained.

BACKGROUND OF THE INVENTION 
Recently, there has developed great interest in compressors which lack 
traditional moving parts. Such compressors are particularly suitable for 
application in refrigeration systems which use non-ozone depleting 
refrigerants as the working medium. One type of such compressor is the 
so-called acoustic compressor whereby a standing acoustic wave is 
generated in a chamber, containing the working medium. The pressure 
variations caused by the acoustic wave are then utilized to pump, and 
therefore compress, the medium. Acoustic compressors are described in 
detail in U.S. Pat. No. 5,020,977 issued to Lucas, the disclosure of which 
is hereby incorporated by reference. 
In the acoustic compressors described in the aforementioned '977 patent, 
the acoustic waveform is generated by a flexible diaphragm mounted at one 
end of the chamber containing the working medium and driven by an 
electromagnetic actuator. The arrangement is thus similar to a 
conventional loudspeaker. The frequency at which the diaphragm is driven 
is chosen in accordance with the physical properties of the working medium 
and the geometry of the chamber so that a standing wave (i.e., one having 
stationary nodes and antinodes) is produced within the chamber. That is, 
the diaphragm is driven at a frequency corresponding to one of the normal 
modes of the system so that a condition of resonance is achieved within 
the chamber. 
Obviously, it is advantageous in a refrigeration system for the acoustic 
compressor to generate pressures as high as possible in order to compress 
the refrigerant. At a given driving frequency, the only way to increase 
the pressure amplitude of the acoustic waveform is to increase the 
amplitude of the diaphragm's back and forth excursion. Such an increase is 
severely constrained, however, by the physical characteristics o both the 
flexible diaphragm and the electromagnetic actuator which permit only a 
limited range of motion at a given frequency. Furthermore, a flexible 
diaphragm can only withstand a limited amount of pressure before it 
physically fails. 
SUMMARY OF THE INVENTION 
It is a primary object of the present invention to provide an acoustic 
compressor capable of generating acoustic waveforms at extremely high 
pressures. Such a compressor is particularly suitable for incorporation 
into a refrigeration system where the acoustic waveform produces 
compression of a refrigerant. 
In accordance with the present invention, a chamber containing the working 
medium is solidly connected to the body of an electromagnetic actuator. 
Within the actuator is an armature which undergoes vibrational motion in 
response to the application of an AC driving signal. The armature is 
mounted within the actuator body by means of armature springs which allow 
the armature to vibrate within the body. The actuator body and chamber are 
also mounted so as to be vibratable, either by means of springs or 
resilient mounting media. Due to conservation of momentum, vibrational 
motion of the armature causes corresponding vibrational motion of the 
chamber containing the working medium. Such vibration, if at a resonant 
frequency determined by the medium and chamber length, produces a standing 
acoustic waveform within the chamber. The resulting pressure variations 
may then be used to pump medium into the chamber through an inlet valve 
and out through an outlet valve. 
The amplitude of the pressure waveform thus generated is dependent upon the 
extent of vibratory displacement undergone by the chamber. The latter can 
be increased by choosing the armature mass and armature springs so that 
the armature vibrates at its resonant frequency. 
Other objects, features, and advantages of the invention will become 
evident in light of the following detailed description considered in 
conjunction with the referenced drawings of a preferred exemplary 
embodiment according to the present invention.

DETAILED DESCRIPTION OF THE INVENTION 
It is well known, of course, that when an oscillating system is driven 
externally at a resonant frequency of the system, the system undergoes 
oscillation at a maximum amplitude limited only by the amount of damping 
(e.g., frictional losses) within the system. Furthermore, at resonance, 
the system absorbs the maximum amount of energy possible from the external 
driving source with no reactive power losses due to back and forth energy 
transfers between the system and the external driver. 
FIG. 1 is a schematic representation of a simple oscillating system 
consisting of masses M1, M2, and M3 connected by springs K1 and K2. (The 
designations also represent the mass and spring constant of each of the 
masses and springs, respectively.) If the mass M1 is caused to oscillate 
at a certain frequency by a driving force, masses M2 and M3 will also 
oscillate owing to the spring connections between the masses. The relative 
amplitudes of oscillation of the masses depends upon the driving 
frequency. If M2 is assumed to be very large, the displacement of mass M3 
will be greatest when its oscillation is equal to the resonant frequency 
at which M3 vibrates with respect to M2. That frequency, F.sub.3, is given 
by: 
##EQU1## 
Displacement of mass M3 is, of course, effected by the displacement of 
mass M2 which transmits force through spring K2. Maximum oscillatory 
displacement of M2 due to a force acting on mass M1 occurs when masses M1 
and M2 oscillate at their resonate frequency. Again, if mass M2 is assumed 
very large, that resonant frequency F.sub.1 is given by: 
##EQU2## 
Therefore, if mass M3 and spring constant K3 are assumed fixed, maximum 
oscillatory displacement of mass M3 occurs when: 
1) the frequency of external driving force applied to mass M1 is equal to 
##EQU3## 
and, 2) mass M1 and spring constant K1 are chosen such that K1/M1=K2/M3. 
The above derivation has assumed mass M2 to be very large relative to 
masses Ml and M3. If such is not the case, the mass M2 must be taken into 
account. Maximum oscillatory displacement of mass M3 at resonant frequency 
##EQU4## 
then occurs when the following equation holds: 
Parameters M1 and K1 correspond to the mass and spring constant 
respectively of the armature, M2 corresponds to the mass of the 
actuator-chamber unit, and M3 and K2 correspond to the dynamic mass and 
spring constant respectively of the fluid in the chamber. The condition 
for maximizing the amplitude of the standing wave set up in the fluid, and 
thus the recurrent fluid pressure variations available for pumping and 
compressing occurs when the frequency F1 of the drive current is made to 
be K1/M1 and the other parameters, K2, M2 and M3, are made to satisfy the 
equation K1/M1=(K2/M3)/(1+M1/M2). 
FIG. 6 shows an embodiment of an acoustic compressor in accordance with the 
present invention. Chamber 14 encloses the working medium 15 (refrigerant 
in the case of a refrigeration system). As the medium is alternately 
compressed and expanded at the end of chamber 14 due to a standing 
acoustic waveform, the medium is pumped into the chamber through inlet 
valve 18 and out through outlet valve 17. Both valves 17 and 18 in this 
embodiment are check valves which allow fluid flow in only one direction. 
The waveform is generated by an electromagnetic driver or actuator 
containing a permanent magnet structure or armature mounted within a body 
or housing by means of springs so as to be vibratable in an axial 
direction. 
FIGS. 3 and 4 show an actuator in greater detail. The armature 25 comprises 
a permanent magnet structure MG1 mounted on an actuator shaft S1 and 
magnetized in an axial direction. The shaft S1 is mounted at its ends 
within a magnetically permeable actuator body H1 which is open at both 
ends along the axis of the shaft S1. The shaft S1 is mounted at each end 
to a leaf spring 10 attached to the body H1. The magnet MG1 is a magnet 
stack which contains one or more coaxially positioned permanent magnets 
and two or more pole plates PP1. A plate PP1 is placed between each magnet 
and at either end of the magnet stack. In case of more than one magnet, 
the magnets are stacked on the shaft S1 in such a way that equal magnetic 
poles face each other. The magnet MG1 in this embodiment is cylindrical 
with a circular cross-section but could be of any arbitrary shape. 
Two or more solenoidal coils C1 are wound on a non magnetic bobbin B1 
mounted within the body H1. The coils C1 are spaced from each other in an 
axial direction so as to form rings on the bobbin B1. The inside diameter 
fits around the magnet structure MG1 with some minimal diametrical 
clearance so as to allow it free axial motion. The outside diameter of the 
bobbin B1 fits snugly inside the cylindrical steel body H1. The coils C1 
are electrically connected with each other in such a way that an 
electrical current flowing through adjacent coils C1 will create opposing 
magnetic fields. When an AC driving signal is applied to the coils Cl, the 
armature 25 undergoes reciprocal motion in accordance with the driving 
signal owing to the forces exerted on the poles of magnet MG1. 
Referring next to FIGS. 4 and 5, there is shown another embodiment of an 
electromagnetic actuator which is of the moving coil type. Two or more 
annular magnet segments MG2 are attached on the inside of an outer 
actuator body H2 made of highly magnetically permeable material such as 
magnetically soft steel. The body H2 is closed at each end with leaf 
springs 10 similar to the embodiment described earlier. The magnets MG2 
are magnetized in a radial direction with adjacent magnets having their 
N-S poles oriented in opposite directions. 
The armature 25 in FIGS. 4 and 5 comprises two or more field coils C2 wound 
on a thin walled magnetic steel cylinder 35. A non-magnetic core 36 
occupies the center of the cylinder 35. The coils C2 are spaced from each 
other with a non-magnetic material 37. An actuator shaft S2 is mounted in 
the center of the core 37 and shaft S2 is mounted within the housing H2 to 
leaf springs 10 attached at each end of the actuator body. The coils C2 
are electrically connected so that an AC input current applied to the 
coils creates opposing magnet fields in each adjacent coil C2 which causes 
the armature 25 to undergo reciprocal motion in accordance with the input 
signal. The body H1 and cylinder 35, provide a magnetic flux path for the 
magnets MG2 which concentrates and directs perpendicularly their fields 
toward the coils C2. 
In either of the embodiments described above, by enclosing the magnets with 
a material having high magnetic permeability (i.e., body H1 or H2), most 
of the magnetic flux is guided through the magnetic gap in which the coils 
C1 or C2 are placed. Only very little flux will leak past this gap, thus 
creating a highly efficient electromagnetic actuator. In the case of the 
first embodiment, all the flux in the pole plates PP1 is guided through 
the coils. In the second embodiment, the cylinder 35 provides a low 
reluctance pathway for the flux produced by the magnets MG2. 
Each of the embodiments of the present invention also provides an actuator 
which presents a low inductance load to the electrical input signal. Since 
the change in input current due to a change in the input voltage signal 
occurs with a time constant proportional to this inductance, reducing this 
inductance results in a more linear relationship between the input signal 
and the force applied to the moving element. The inductance is reduced by 
electrically connecting adjacent coils in series so as to create opposing 
magnetic fields. An inductive coupling between adjacent coils occurs due 
to a linkage of magnetic fields created when a current runs through the 
coils. When the direction of the current in two adjacent coils is the same 
the resulting magnetic fields are added and the total inductance of the 
combination is double the inductance of a single coil. When the currents 
run in opposite directions, the induced magnetic fields oppose each other 
which reduces the total inductance. Since the electrical time constant is 
proportional to the total inductance, it follows that the time constant is 
also thereby reduced. This effective reduction in total inductance is 
enhanced the better is the flux linkage between the coils. The 
above-described embodiments achieve a high degree of flux linkage between 
adjacent coils by providing a low reluctance pathway inside the coils. 
The embodiments described above also provide an actuator where the force on 
the moving armature (i.e., either the coil or the magnet) due to a given 
input current is independent of the position of the armature along its 
stroke path. This is achieved in the first embodiment by providing coils 
C1 adjacent to each pole of the magnets MG1 making up the armature 25. 
Similarly, in the second embodiment, an annular magnet MG2 is provided 
adjacent each of the coils C2. By properly manipulating the dimensional 
relations of the coils and the pole plates the magnetic flux flowing 
through the coils can be made constant over the length of the moving 
element's stroke. 
Referring again to FIG. 6, chamber 14 is solidly connected to body 4 of the 
electromagnetic actuator or driver. The cylindrical driver contains a 
magnetic armature 25 which consists of a magnet 5, two pole pieces 6a-b, 
two spacers 7, two washers 8 and two screws 9. The armature is 
concentrically placed in the driver body 4 and supported by two leaf-type 
springs 10. The springs 10 are clamped to the driver body 4 by means of 
the clamping rings 11 and the screws 12. Two coils 13a-b are placed 
between the pole pieces 6 and the driver body 4 leaving a small clearance 
between the coil and the pole piece. The coils 13 are solidly adhered to 
the driver body 4. Due to the axial direction of the magnetization of the 
magnet 5, the pole pieces 6 exhibit opposite magnetic polarity. Hence the 
magnetic flux from pole piece 6a flows through coil 13a to the soft 
magnetic driver body 4 through the coil 13b and back to the other pole 
piece 6b, thus creating opposing magnetic fields in the coils 13. The 
coils 13 are electrically connected in such a way that the current running 
in one coil is in opposite direction from the current running in the other 
coil, thus creating a low inductance coil pair and generating oppositely 
directed magnetic fields. When an electrical current flows through the 
coils 13, a force is exerted in the armature in an axial direction and 
proportional to the amount of current. When an alternating current is 
applied to the coils 13, the resulting alternating force produces 
reciprocating motion of the armature. 
If the entire compressor shown in FIG. 6 were assumed to be suspended in 
free space, vibrational motion of the armature as described above would 
produce similar motion in opposite phase of the body 4 and chamber 14 due 
to conservation of momentum (i.e., due to the reaction force acting on the 
coils 13 produced by the magnetic field.) This result can be approximated 
by resiliently mounting the compressor on springs or a similar structure 
allowing some freedom of movement. FIG. 6 shows that the compressor is 
mounted on a frame 20 by means of leaf springs 22 extending from chamber 
14 so as to allow some vibrational motion in the axial direction. 
The components shown in FIG. 6 can now be seen to be approximately 
represented by the oscillating system of FIG. 1. The armature is 
equivalent to mass M1, leaf springs 10 are equivalent to spring K1, the 
chamber 14 and remainder of the actuator including the body 4 are 
equivalent to mass M2 and the resonant frequency K2/M3 is equal to one of 
the resonant frequencies of the system which is dependent upon the 
physical properties of the fluid 15 and the geometry of chamber 14. FIG. 7 
shows a representation of the chamber 14 where the medium within is 
modeled by a mass 1 connected to the chamber walls by springs 2 which 
corresponds to the fundamental vibrational mode of the actual system. 
FIG. 8 is a graph of fluid pressure versus position along the length of the 
chamber 14 for two cases corresponding to the limits of the chamber's 
vibratory motion. The broken curve illustrates the situation when the 
chamber has moved to the left while the solid curve shows the pressure 
distribution when the chamber has moved to the right. The result is a 
standing wave with a node in the middle and antinodes at both ends of the 
chamber. Other resonant modes are, of course, possible which also produce 
standing wave patterns when the chamber 14 is oscillated. The necessary 
condition is that the driving frequency corresponds to a wavelength 
.lambda. within the medium such that the chamber length equals n.lambda./2 
where n is an odd integer. (Note that this is different from the 
compressors described in the '977 patent which use a single moving 
diaphragm. There, the chamber length must equal n.lambda./4 where n is an 
integer in order to produce resonance.) 
By using the electromagnetic actuator to produce vibratory motion of the 
chamber containing the working fluid, it is possible to produce much 
greater pressures within the chamber than with a conventional single 
diaphragm. This is because the chamber may be displaced to a greater 
extent during its vibratory cycle than can a diaphragm. Furthermore, 
energy is applied to both ends of the chamber simultaneously since a 
rarefaction occurs at one end while a compression occurs at the other. 
Energy transfer is also made more efficient because the reaction forces 
produced by the actuator are not absorbed in the frame but in the chamber 
itself. Maximum vibratory displacement of the chamber, and therefore, 
maximum amplitude of the pressure waveform within, occurs when the 
resonant frequency of the armature 25 and leaf springs 10 combination 
equals a resonant frequency of the chamber 14 containing the medium. 
Although the invention has been described in conjunction with the foregoing 
specific embodiment, many alternatives, variations, and modifications will 
be apparent to those of ordinary skill in the art. Those alternatives, 
variations, and modifications are intended to fall within the scope of the 
following appended claims.