U.S. patent application number 11/759295 was filed with the patent office on 2007-12-13 for capacity control of a compressor.
This patent application is currently assigned to TECUMSEH PRODUCTS COMPANY. Invention is credited to Dan M. Manole, Phillip A. Tomell.
Application Number | 20070286751 11/759295 |
Document ID | / |
Family ID | 38822208 |
Filed Date | 2007-12-13 |
United States Patent
Application |
20070286751 |
Kind Code |
A1 |
Manole; Dan M. ; et
al. |
December 13, 2007 |
CAPACITY CONTROL OF A COMPRESSOR
Abstract
A linear compressor that is operated at a frequency greater than
the natural frequency of the spring-mass system of the compressor.
Operating the compressor at such a frequency can increase the
output of the compressor. In one embodiment, the linear compressor
includes a cylinder block having a cylinder bore, a piston
positioned within the cylinder bore, first and second springs for
positioning the piston where the piston and the first and second
springs comprise the spring-mass system, and an armature operably
engaged with the piston to drive the piston at a frequency greater
than the natural frequency of the spring-mass system. The linear
compressor can also include a controller which monitors the
instantaneous natural frequency of the spring-mass system and
modulates the frequency of the current passing through the armature
such that it exceeds the natural frequency of the spring-mass
system.
Inventors: |
Manole; Dan M.; (Tecumseh,
MI) ; Tomell; Phillip A.; (Adrian, MI) |
Correspondence
Address: |
BAKER & DANIELS LLP;111 E. WAYNE STREET
SUITE 800
FORT WAYNE
IN
46802
US
|
Assignee: |
TECUMSEH PRODUCTS COMPANY
100 East Patterson Street
Tecumseh
MI
49286
|
Family ID: |
38822208 |
Appl. No.: |
11/759295 |
Filed: |
June 7, 2007 |
Related U.S. Patent Documents
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Application
Number |
Filing Date |
Patent Number |
|
|
60812890 |
Jun 12, 2006 |
|
|
|
Current U.S.
Class: |
417/417 |
Current CPC
Class: |
F04B 2201/0806 20130101;
F04B 2203/0404 20130101; F04B 35/045 20130101 |
Class at
Publication: |
417/417 |
International
Class: |
F04B 35/04 20060101
F04B035/04 |
Claims
1. A linear compressor, comprising: a cylinder block having a
cylinder bore; a piston, wherein at least a portion of said piston
is positioned within said cylinder bore; a first spring for
positioning said piston, said piston and said first spring
comprising a spring-mass system having a natural frequency; an
armature operably engaged with said piston to drive said piston at
a driving frequency; and a controller for increasing said driving
frequency above said natural frequency.
2. The linear compressor of claim 1, further comprising a sensor
for measuring the frequency of electrical current flowing through
said armature, said sensor in communication with said
controller.
3. The linear compressor of claim 1, further comprising a sensor
for measuring one of the temperature and the pressure of a
refrigerant compressed by said compressor.
4. A method for operating a linear compressor, comprising the steps
of: providing a cylinder block having a cylinder bore, a piston
positioned within said cylinder bore, and at least one spring for
positioning said piston, said piston and said at least one spring
comprising a spring-mass system having a natural frequency;
applying a driving force to said piston to drive said piston within
said cylinder bore at a driving frequency; and increasing said
driving frequency above said natural frequency, whereby said
increasing step increases the capacity of said linear
compressor.
5. The method of claim 2, further comprising the step of monitoring
said natural frequency.
Description
BACKGROUND OF THE INVENTION
[0001] 1. Field of the Invention
[0002] The present invention relates to compressors, and more
particularly, to the capacity control of linear compressors.
[0003] 2. Description of the Related Art
[0004] Compressors can include a piston which is reciprocated
within a cylinder bore to compress refrigerant, for example, in the
cylinder bore. The compressor can further include a spring, or
springs, which bias the piston into position. In some linear
compressors, the piston is positioned intermediate two springs
which hold the piston in a substantially stationary position until
the piston is moved by an electromagnetic armature or motor, for
example. The piston and springs comprise a spring-mass system
having a natural, or resonant, frequency, as known in the art. If
the piston is driven, via the armature or motor, at the natural
frequency of the spring-mass system, the spring-mass system will
resonate. Driving the piston of the compressor at, or very close
to, the natural frequency of the system allows the compressor to
operate more efficiently. In effect, when the spring-mass system is
driven at, or close to, its natural frequency, the driving force
has less inertial forces in the system to overcome.
[0005] In view of the above, previous compressors were typically
operated at the natural frequency of their spring-mass systems. To
increase or decrease the capacity of these compressors, the
displacement, or stroke, of the piston was adjusted to change the
output of the compressor. For example, if a greater capacity was
needed, the stroke of the piston was increased to draw in,
compress, and discharge a larger quantity of refrigerant per
stroke. To increase the stroke of the piston, the magnitude of the
current flowing through the armature was increased, thereby causing
a greater displacement between the piston and the armature.
However, modulating the capacity of the compressor in this way has
some limitations. For example, increasing the magnitude of the
current flowing through the armature can increase the resistance
losses in the armature windings, thereby reducing the efficiency of
the compressor. Further, large displacements of the piston draws
large quantities of refrigerant into the cylinder bore which may
bog down or overpower the compressor.
[0006] Previously, as discussed above, it was desirable to operate
linear compressors at the natural frequency of their spring-mass
system. However, owing to changes in the parameters of the
refrigerant in the cylinder bore, the natural frequency of the
spring-mass system can change throughout the operation of the
compressor. More specifically, when the refrigerant is compressed
by the piston in the cylinder bore, the refrigerant gas acts as an
elastic spring force against the piston. The magnitude of this
elastic force depends on, among other things, the fluid being
compressed and its density, pressure, and temperature. As known in
the art, the magnitude of the spring force from the refrigerant gas
affects the natural frequency of the spring-mass system, and, when
the parameters of the refrigerant change, the natural frequency of
the spring-mass system typically changes as well. In order to
determine the natural frequency of the spring-mass system at any
instant during the operation of the compressor, a parameter, or
parameters, of the refrigerant and/or refrigeration system can be
monitored. For example, it was known to monitor temperature of the
refrigerant and/or the voltage drop across the armature driving the
piston of the compressor. In view of the information obtained from
monitoring these parameters, the frequency of the driving force
acting on the piston was altered to match the instantaneous natural
frequency of the system.
[0007] In effect, some previous compressors actively monitored the
natural frequency of the spring-mass system and corrected the
frequency of the driving force to match the natural frequency of
the system. However, when these compressors were required to
produce a greater output of compressed refrigerant, their output
was limited to that generated at the natural frequency of the
compressor. As a result, as discussed above, these compressors were
sometimes unable to keep up with the demands of the refrigeration
system. To accommodate a potentially greater demand, a compressor
having a larger capacity was typically used. However, these
larger-capacity compressors are typically more expensive and may
become less efficient when lower demands of the compressor are
required. What is needed is an improvement over the foregoing.
SUMMARY OF THE INVENTION
[0008] The present invention includes a linear compressor that is
operated at a frequency greater than the natural frequency of the
spring-mass system of the compressor. Operating the compressor at
such a frequency can increase the output of the compressor. In one
embodiment, the linear compressor includes a cylinder block having
a cylinder bore, a piston positioned within the cylinder bore,
first and second springs for positioning the piston where the
piston and the first and second springs comprise a spring-mass
system having a natural frequency, and an armature operably engaged
with the piston to drive the piston at a frequency greater than the
natural frequency of the spring-mass system.
[0009] In another embodiment, the linear compressor includes a
controller which monitors the instantaneous natural frequency of
the spring-mass system and modulates the frequency of the current
passing through the armature. As discussed above, the natural
frequency of the spring-mass system can change as a result of
fluctuations in the temperature and/or pressure of the refrigerant
in the cylinder bore. In this embodiment, a parameter of the
refrigerant in the refrigerant circuit, such as the pressure and/or
temperature of the refrigerant, for example, or the electrical
power transmitted to the armature, such as the voltage and/or
current, for example, is monitored by the controller. In view of
the information obtained from monitoring these parameters, the
controller can determine the instantaneous natural frequency of the
spring-mass system and evaluate whether the frequency of the
current being transmitted to the armature is greater than the
instantaneous natural frequency of the system. If necessary, the
controller can increase the frequency of the current such that it
exceeds the natural frequency of the spring-mass system, or, even
if the driving frequency is already greater than the natural
frequency, it can increase the driving frequency to increase the
output of the compressor to meet the demands of the refrigeration
system.
BRIEF DESCRIPTION OF THE DRAWINGS
[0010] The above-mentioned and other features and objects of this
invention will become more apparent and the invention itself will
be better understood by reference to the following description of
an embodiment of the invention taken in conjunction with the
accompanying drawings, wherein:
[0011] FIG. 1 is a schematic of a typical refrigeration circuit
including a compressor and a controller for operating the
compressor;
[0012] FIG. 2 is a partial cut-away view of a linear compressor in
accordance with an embodiment of the present invention;
[0013] FIG. 3 is a cross-sectional perspective view of a first
alternative embodiment linear compressor;
[0014] FIG. 4 is an exploded cross-sectional view of a second
alternative embodiment linear compressor;
[0015] FIG. 5 is a schematic representing the spring-mass system of
the compressor of FIG. 2; and
[0016] FIG. 6 is a graph charting the cooling capacity of a linear
compressor with respect to the current flowing through the armature
of the compressor.
[0017] Corresponding reference characters indicate corresponding
parts throughout the several views. Although the exemplifications
set out herein illustrate embodiments of the invention, the
embodiments disclosed below are not intended to be exhaustive or to
be construed as limiting the scope of the invention to the precise
form disclosed.
DETAILED DESCRIPTION
[0018] Referring to FIG. 1, typical refrigeration system 10
includes, in serial order, compressor 12, condenser 14, expansion
device 16, and evaporator 18 connected in series by fluid conduits.
As is well known in the art, compressor 12 draws a refrigerant or
working fluid through compressor inlet 11, compresses the
refrigerant, and expels the compressed refrigerant through
compressor outlet 13. The refrigerant expelled from compressor 12
is communicated into condenser 14 where thermal energy of the
refrigerant is dissipated. Subsequently, the cooled, compressed
refrigerant is communicated to expansion device 16 where it is
decompressed. The cooled, low-pressure refrigerant is then
communicated to evaporator 18 where the refrigerant in evaporator
18 draws heat from an environment surrounding the evaporator.
Subsequently, the refrigerant exits evaporator 18 and is
communicated to compressor 12 and the cycle described above is
repeated.
[0019] Referring to FIG. 2, compressor 12, in the present
embodiment, is a dual-cylinder linear compressor having two
axially-driven compressor mechanisms 48 mounted therein. Compressor
12 further includes housing 42 having interior cavity 44 and end
caps 46 on opposite ends thereof which also define cavity 44.
Generally, in operation, refrigerant is drawn into compressor 12
through suction inlet 11 and suction manifold 45, compressed by
compressor mechanisms 48, and is then discharged into discharge
muffler chamber 51 through discharge valves 55. Referring to FIG.
4, which illustrates an alternative embodiment of a linear
compressor, each compressor mechanism 48 can include gasket 61,
suction valve 59, valve plate 53, and discharge valve 55 for
controlling the flow of suction refrigerant into, and the flow of
discharge refrigerant out of, the compression cylinder of
compressor mechanism 48. Thereafter, the compressed refrigerant is
discharged from compressor 12 through discharge outlet 13.
[0020] Each compressor mechanism 48 includes a cylinder block 50
having cylinder bore 52 therein, a piston 54 positioned within
cylinder bore 52, an armature 56 mounted to one end of piston 54,
and a permanent magnet 58 positioned within end cap 46. In
operation, piston 54 is reciprocatingly driven within cylinder bore
52 by the interaction of armature 56 and permanent magnet 58. More
particularly, armature 56 is energized by an electrical source
which conducts electricity to armature 56 through terminal cluster
60 and spring 62 positioned intermediate cylinder block 50 and
armature 56. Armature 56 includes a series of copper windings, or
coils, which are, in this embodiment, arranged in a cylindrical
configuration. The cylindrical configuration of armature 56 is
sized and configured to fit in gap 66 defined between permanent
magnet 58 and end cap 46 so that relative movement of armature 56
therebetween is possible. Owing to a magnetic field created by
permanent magnet 58, armature 56, when energized, is motivated to
move axially along axis 64.
[0021] Permanent magnet 58, as is known in the art, contains two
poles of opposite polarity which create the above-mentioned
magnetic field. The magnetic field of permanent magnet 58 radiates
through bottom 47 of end cap 46, through side walls 49 of end cap
46, and then to the other pole of permanent magnet 58 trough
annular air gap 66 between end cap 46 and permanent magnet 58.
Stated in another way, the magnetic field extends through gap 66 in
a radial direction, i.e., a direction substantially perpendicular
to axis 64. As the coils of armature 56 are positioned in gap 66,
the magnetic field crosses the coils and interacts with the current
flowing through the coils to generate Lorenz forces that will move
armature 56 in a direction perpendicular to the electrical current
and the magnetic field, i.e., along axis 64. By alternating the
current polarity, the direction of the axial force acting on
armature 56 can be changed to reciprocate armature 56, and piston
54 attached thereto, along axis 64.
[0022] In the present embodiment, the armature is mounted on the
piston and the stationary permanent magnet is mounted in the
housing. However, in other embodiments, the permanent magnet may be
mounted on the reciprocating piston and the armature may be
stationary within the compressor.
[0023] As discussed above, compressor mechanism 48 includes spring
62 positioned between armature 56 and cylinder block 50. Compressor
mechanism 48 further includes second spring 68 positioned between
armature 56 and permanent magnet 58 positioned in end cap 46.
Springs 62 and 68 act to hold armature 56, and piston 54 mounted
thereto, in a substantially stationary position until the coils of
armature 56 are energized. Also, spring 68 completes the electrical
circuit between terminal cluster 60, spring 62 and armature 56, as
described above. Once energized, one of springs 62 and 68,
depending on the polarity of the current, is compressed by the
Lorenz forces acting on armature 56 placing piston 54 in one of a
top-dead-center (TDC) position or a bottom-dead-center (BDC)
position. The TDC and BDC positions define the limits of the stroke
of piston 54 within cylinder bore 52, however, the distance between
the TDC and BDC positions is dependent upon the root mean square
average value (RMS), or magnitude, of the current passing through
the armature. For example, the TDC and BDC positions are further
apart from each other when the RMS of the current passing through
the armature is increased, and, as a result, the TDC and BDC
positions define a longer stroke of the piston and a potentially
larger output of refrigerant.
[0024] Referring to FIG. 5, piston 54 and springs 62 and 68
approximate a spring-mass system. Generally, a spring-mass system
represents a harmonic system that satisfies the second order
differential equation: x=A sin .omega..sub.0t+B cos .omega..sub.0t,
where x represents the displacement of piston 54 and .omega..sub.0
represents the circular natural frequency, where
.omega..sub.0=(k/m) 0.5 and is typically measured in radians per
second. The constant k represents the spring constant of the
spring-mass system, including, in the present embodiment, the
spring constants of springs 62 and 68, and the constant m
represents, in the present embodiment, the combined mass of piston
54, armature 56, and 1/3 of the mass of springs 62 and 68. In the
foregoing equation, A and B are determined by an initial driving
input into the system. The natural frequency of this spring mass
system is determined by the following equation:
f=.omega..sub.0/(2.pi.)=((k/m) 0.5)/(2.pi.). When the spring-mass
system is driven by a force having a frequency matching, or nearly
matching, the natural frequency of the spring-mass system, the
system will resonate. In resonance, the piston of the spring-mass
system will have less inertia in the system to overcome and, as a
result, less power is required to operate the compressor.
Accordingly, compressor manufacturers previously designed their
linear compressors to operate at the natural frequency of the
linear compressor's spring-mass sytem in order to utilize this
phenomenon.
[0025] To increase the output of these previous compressors, the
RMS of the current flowing through the armature is increased while
the frequency of the current is held at the natural frequency.
Increasing the RMS of the current causes the piston and armature
assembly to displace through a greater distance, thereby increasing
the stroke and output of the compressor. However, the stroke of the
piston is ultimately limited by the length of the cylinder bore
and, thus, some adjustments to the compressor capacity may not be
possible. Further, by increasing the stroke of the piston, a
greater quantity of refrigerant enters into the cylinder bore per
stroke which may be difficult for the compressor to compress,
thereby bogging down the compressor. In addition, increasing the
RMS of the current flowing through the armature can increase the
resistance losses in the armature windings, thereby reducing the
efficiency of the compressor, as illustrated in the following
example. Referring to FIG. 6, the operating condition of a
compressor is represented by point 1. Notably, an increase in the
RMS of the current, I.sub.DC, increases the cooling capacity, Q, of
the compressor as long as the operating condition of the
compressor, represented by point 1, is to the left of line 70.
However, for operating conditions located to the right of line 70,
such as point 2, an increase in the RMS of the current will
actually reduce the cooling capacity of the compressor owing to
losses in the armature. In addition, the cooling capacity of the
compressor can be controlled by adjusting the duty cycle, D, of the
current through the armature. The duty cycle is the ratio of the
pulse duration of the current to the pulse period, i.e., the ratio
of the duration of when the windings are energized divided by the
time between the beginning of one energization and the next.
Referring to FIG. 6, the duty cycle of the armature current can be
increased to the point where the operating condition of the
compressor is to the right of line 70, such as point 3, where any
additional increase in the duty cycle actually decreases the
cooling capacity of the compressor. To prevent such occurrences in
these previous compressors, larger-capacity compressors may be
necessary. Larger-capacity compressors are typically more expensive
and less efficient at lower capacities.
[0026] While a linear compressor may generally approximate the
harmonic spring-mass system described above, this approximation is
somewhat simplified. For example, the above equations do not
account for damping, or losses, in the system. Most spring-mass
systems are at least somewhat damped, i.e., they have losses in the
system which dissipate energy and cause the motion of the piston to
gradually decay. Further, although the above equations account for
an initial input, they do not account for a continuous driving
force. An equation mathematically representing a spring-mass system
which accounts for system damping and a continuous driving force is
d.sup.2x/dt.sup.2+(b/m)dx/dt+(k/m)x=A.sub.0 cos(.omega.t), where b
represents the damping coefficient of the system and .omega.
represents the circular frequency of the driving force applied to
the mass.
[0027] As indicated above, the spring constant k for the
spring-mass system is mostly defined by the spring constants of
springs 62 and 68, i.e., k.sub.1 and k.sub.2, respectively. As
known in the art, the spring constants of linear springs, for
example, are substantially constant, although they may change
slightly throughout their use. Referring to FIG. 5, in addition to
the spring forces applied to piston 54 by springs 62 and 68, the
refrigerant within cylinder bore 52 can apply an additional spring
force acting on piston 54. More particularly, as the refrigerant in
cylinder bore 52 is compressed by piston 54, the pressure of the
refrigerant in cylinder bore 52 increases and the pressurized
refrigerant exerts a force against piston 54. The refrigerant
substantially acts like an elastic spring where the spring
stiffness of the compressed refrigerant can be represented by a
spring constant, i.e., k.sub.3. In view of this, in the present
embodiment, the spring stiffness, k, of the spring-mass system
equals k.sub.1+k.sub.2+k.sub.3. However, owing to changes in
pressure and temperature of the refrigerant, for example, the
spring stiffness of the refrigerant, k.sub.3, may change throughout
the operation of the compressor. For example, a change in either
the temperature or pressure of the refrigerant may increase the
stiffness, k.sub.3, of the refrigerant whereas a decrease in the
temperature or pressure may decrease the stiffness. As the spring
stiffness, k, of the spring-mass system affects the natural
frequency of the system, when k.sub.3 changes, the natural
frequency of the spring-mass system changes.
[0028] Unlike previous compressors, compressors embodying the
present invention are designed such that the frequency, .omega., of
the driving force acting on the spring-mass system can be increased
above the natural frequency of the spring-mass system. In the
illustrated embodiment, the frequency of the current passing
through armature 56 determines the frequency of the driving force
acting on the spring-mass system. In this embodiment, the frequency
of the current substantially equals the frequency of the driving
force. Accordingly, to increase the frequency of the driving force
acting on the spring-mass system, for example, the frequency of the
current passing through armature 56 is increased. Increasing the
frequency of the driving force acting on the spring-mass system can
increase the strokes per minute of piston 54 within cylinder bore
52, thereby potentially increasing the output of the compressor. In
one embodiment, the frequency of the current is increased without
increasing the magnitude of the current, i.e., without increasing
the stroke length of piston 54. In this embodiment, the potential
disadvantages described above with respect to previous compressors
can be avoided. However, in other embodiments, in addition to
increasing or decreasing the output of the compressor through
frequency modulation, the output of the compressor can also be
increased or decreased by modulating the magnitude of the current
and, thus, the stroke length of the piston.
[0029] As discussed above, the natural frequency of a spring-mass
system can change throughout the operation of a compressor owing to
changes in the temperature and/or pressure of the refrigerant being
compressed in the cylinder bore of the compressor, for example. In
one embodiment, the range of potential natural frequencies can be
determined before the compressor is placed into service and the
minimum frequency of the driving force can be set above the maximum
potential natural frequency of the spring-mass system. In this
embodiment, the natural frequency of the driving force can be
established without continuously monitoring the parameters of the
refrigerant being compressed. While this embodiment is a
contemplated embodiment of the present invention, other embodiments
are envisioned where the parameters of the refrigerant being
compressed, or other parameters of the refrigeration system, are
monitored and the frequency of the driving force is adjusted
accordingly.
[0030] In one embodiment, the compressor includes a controller,
such as controller 26 (FIG. 1), which monitors at least one system
parameter and, in view of the information obtained from monitoring
this parameter, makes a running correction to the frequency of the
current passing through electrical wires 28 and armature 56. In one
embodiment, referring to FIG. 1, temperature sensors 20 and 22 are
placed in communication with the flow of refrigerant entering into
and flowing out of compressor 12. In one embodiment, the controller
is programmed with a first table of data that correlates the
temperature of the refrigerant at one or both of these locations
with a natural frequency of the spring-mass system. This table of
data can be established empirically or via equations which
calculate the spring constant k.sub.3 of the refrigerant in the
cylinder bore and, ultimately, the natural frequency of the
system.
[0031] In a further embodiment, the controller can monitor a
parameter of the electrical power provided to armature 56
including, for example, the current flowing through the armature or
the voltage measured across the armature during operation. In one
embodiment, the controller can be programmed with a second table of
data which correlates the parameters of the armature current and/or
voltage with the instantaneous natural frequency of the system. The
controller can be programmed to compare the data in the first table
and the data in the second table and make a running correction to
the current flowing through the armature. The data contained in the
second table can be derived from equations which associate the
operating parameters of the compressor with the natural frequency
of the system. In one embodiment, the equation,
I.sub.max=f*W.sub.cycle/(D*U.sub.max) can be utilized, where
W.sub.cycle represents the work performed by the compressor per
cycle of the compressor, where D represents the duty cycle of the
current passing through the armature, where U.sub.max represents
the voltage drop across the armature, and where I.sub.max
represents the current passing through the armature. In one
embodiment, the controller can include a frequency converter for
converting the frequency of the current flowing into the armature
to another frequency. The frequency converter can include
electromechanical and/or solid state components, as known in the
art.
[0032] In another embodiment, the instantaneous natural frequency
of the spring-mass system can be determined by measuring the
vibrations of the compressor. In one embodiment, an accelerometer
can be affixed to the compressor housing and/or cylinder block, for
example, to measure the vibrations produced by the spring-mass
system. As known in the art, a spring-mass system produces
different vibrations when the system is driven at its natural
frequency as compared to when the system is driven above or below
its natural frequency. The accelerometer can be placed in
communication with the controller where the controller evaluates
whether the spring-mass system is being driven at its natural
frequency and makes any necessary adjustments to the frequency
and/or magnitude of the current passing through the armature.
[0033] As discussed above, the capacity of compressors utilizing an
embodiment of the present invention can be adjusted via adjustments
to the current flowing through the armature. If the frequency of
the current flowing through the armature is increased, the piston
will typically be cycled through more strokes per minute. Likewise,
if the frequency of the current is decreased, then the piston will
be cycled through less strokes per minute. In view of this, a
compressor which has a stroke that is close to its physical
boundaries in the cylinder bore can be used and still provide
capacity modulation for the refrigeration system. As a result, a
smaller, less-expensive compressor can be used.
[0034] Although the advantages of operating the above-described
compressors above the natural frequency of their spring-mass
systems have been outlined herein, the compressors of the present
invention are nonetheless capable of being operated at or below the
natural frequency of their spring-mass systems. These circumstances
typically arise when the compressor is being cycled on or off
and/or when the demands of the refrigeration circuit drop and a
lower output of the refrigeration system is required.
[0035] While this invention has been described as having an
exemplary design, the present invention may be further modified
within the spirit and scope of this disclosure. This application is
therefore intended to cover any variations, uses, or adaptations of
the invention using its general principles. Further, this
application is intended to cover such departures from the present
disclosure as come within known or customary practice in the art to
which this invention pertains.
* * * * *