Note: Descriptions are shown in the official language in which they were submitted.
CA 02591359 2007-06-12
Dan M. Manole
Phillip A. Tomell
CAPACITY CONTROL OF A COMPRESSOR
BACKGROUND OF THE INVENTION
I. Field of the Invention
[0001] The present invention relates to compressors, and more particularly, to
the capacity
control of linear compressors.
2. Description of the Related Art
[0002]_ 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 annature 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.
100031 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.
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[00041 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.
[00051 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
[0006J 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
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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.
[0007] 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
[0008] 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:
[0009] Fig. 1 is a schematic of a typical refrigeration circuit including a
compressor and a
controller for operating the compressor;
[0000] Fig. 2 is a partial cut-away view of a linear compressor in accordance
with an
embodiment of the present invention;
[0010] Fig. 3 is a cross-sectional perspective view of a first alternative
embodiment linear
compressor;
[0011] Fig. 4 is an exploded cross-sectional view of a second alternative
embodiment
linear compressor;
[0012] Fig. 5 is a schematic representing the spring-mass system of the
compressor of
Fig. 2; and
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[0013] 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.
[0013] 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
[0014) 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.
[0015] 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 I 1 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.
[0016] 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
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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 tenminal 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 defmed 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.
[0017] 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.
[0018] 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.
[0019] 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 a nd 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
CA 02591359 2007-06-12
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.
[0020] Refenring 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 coQt + B cos wot, where x
represents the
displacement of piston 54 and wo represents the circular natural frequency,
where wo =
(k/m)10.5 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, anmature 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= tuo
l(27c) _((k/m)~0.5)l(2n). 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.
[0021] 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 anmature
windings, thereby
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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, IDc, 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
occun:ences in these previous compressors, larger-capacity compressors may be
necessary.
Larger-capacity compressors are typically more expensive and less efficient at
lower
capacities.
[0022] 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 d2x/dt2 +(b/m)dx/dt +(k/m)x = Ao cos(wt), where b
represents the
damping coefficient of the system and w represents the circular frequency of
the driving force
applied to the mass.
[0023] 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., kl and k2,
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
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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., k3. In
view of this, in the
present embodiment, the spring stiffness, k, of the spring-mass system equals
k, + k2 + k3.
However, owing to changes in pressure and temperature of the refrigerant, for
example, the
spring stiffness of the refrigerant, k3, 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, k3, 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 k3 changes, the natural frequency of the spring-
mass system
changes.
[00241 Unlike previous compressors, compressors embodying the present
invention are
designed such that the frequency, w, 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.
[0025) 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
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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.
[0026] 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 k3 of the refrigerant in the cylinder bore and, ultimately,
the natural frequency
of the system.
[0027) 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,,,. = f* W.,.1e /(D *U.) can be utilized, where
Wy,ie 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 Um. represents
the voltage
drop across the armature, and where It. represents the current passing through
the annature.
In one embodiment, the controller can include a frequency converter for
converting the
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frequency of the current flowing into the armature to another frequency. The
frequency
converter can include electro-mechanical and/or solid state components, as
known in the art.
[0028] 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.
[0029] 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.
[0030] 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.
[0031] 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.
