Note: Descriptions are shown in the official language in which they were submitted.
1~96920
This invention relates to reciprocating electric motors
of the type in which a magnetically permeable armature is disposed
to axially reciproeate within the air-gap of a fixed electromag-
netic circuit.
This application is a divisional application of copen-
ding application No. 235,045 filed Sept. 9, 1975.
In the art of moving-iron linear-motor compressors,
much effort has been expended without having achieved significant
commereial suceess (see P.W. Curwen, "Recent Developments of Oil-
Free Linear-Motor Resonant-Piston Compressors", ASME publication
69-FE-36, June, 1969). The linear-motor compressor disclosed
herein has been subjected to extensive laboratory testing and the
design parameters have been verified through the use of iterative
eomputer programming teehniques, and therefore, the requirements
of a commercially viable product are believed to have at last been ;
aehieved.
The present invention provides a moving-iron linear-
motor compressor that is easy and eeonomieal to assemble.
The present invention also provides a magnetie eore for
a moving-iron, linear-motor eompressor whieh may be more easily
assembled and whieh has less waste material than a eore of eon-
ventionally staeked laminations. It is a related objeet to pro-
vide an economieal method for manufacturing said magnetic core.
The present invention also provides a magnetic core
for a moving-iron, linear-motor compressor which is incompressible
in the direction of tightening of the compressor tie bolts in
order to insure that the tie bolts continue to exert the reten-
tion stresses to which they are initially set during assembly to
thereby maintain proper alignment of the core, armature and frame
of the compressor.
-- 1 --
.
1~69ZC~
The present invention also provides a moving-iron,
linear-motor compressor which is small in size and which may be
easily packaged to suit a wide variety of applications.
The present invention further provides a moving-iron,
linear-motor compressor in which the twisting forces imparted on
the reciprocating armature are reduced substantially to zero.
The` present application is an improvement on the linear
compressor disclosed in U.S. Patent No. 3,947,155, Michael K.
Bidol.
To reduce the twisting forces imparted on the recipro-
cating armature of the compressor motor, the present invention
provides a moving-iron, linear-motor compressor wherein the means
for returning the armature at the end of its magnetically powered
stroke comprises a pair of linear internested coil springs each
having substantially zero pitch.
The present invention further provides a moving-iron,
linear-motor compressor having a magnetic core which includes a
pair of abutting convoluted loops of magnetic strip r.laterial
spirally would on a plane perpendicular to the compressor axis of
reciprocation with the magnetic core air-gap formed in the area
of abutment of the two loops.
In copending applicat~on No. 235,045 there is provided
in an electromagnetic compressor of the type including a gas
pump having a pumping chamber, a magnetic circuit including means
fixedly attached to said pump and naving an air-gap, and recip-
rocating means attached to a pumping member and disposed to
axially reciprocate within said air-gap, means for magnetically
activating said magnetic circuit to drive said reciprocating
means in a first direction, and return means operatively con-
nected to said reciprocating means to drive said reciproat-
;92(~
ing means in a direction opposite to said first direction, saidreturn means including a a pair of parallel coil springs, means
clamping one end of each said springs in fixed relation to said
magnetic circuit, and means clamping the other end of each said
springs to said reciprocating means, the improvement wherein said
coil springs comprise a pair of linear coil springs having
substantially zero pitch.
According to one aspect of the present invention
there is provided an electromagnetic circuit of the type in which
a magneticall~ permeable armature is disposed to axially
reciprocate within an air-gap defined by a pair of spaced-apart,
coaxial poles on a fixed magnetic core, said air-gap comprising
means providing a first space between said poles, said space
having a circular cross section perpendicular to the axis of
reciprocation of said armature, and means providing opposing pairs
of exposed coplanar gap surfaces, the planes of said surfaces
being displaced on opposite sides of said axis of reciprocation
such that the maximum displacement between said pairs of surfaces
is less than the maximum diameter of said first space, the space
between said pairs of surfaces being devoid of core material,
the improvement wherein said maximum displacement between said
pairs of surfaces is greater than 16.7% of the cross-sectional
area of said poles as measured on a plane-parallel to said
pairs of surfaces.
According to another aspect of the invention there
is provided an electromagnetic compressor of the type including
a gas pump having a pumping chamber with a central axis, a
magnetic circuit having an air-gap co-axial with said pumping
chamber, and means fixedly attaching said magnetic circuit to
said pump, said attaching means being oriented to exert retention
stresses on said magnetic circuit in a direction parallel to said
central axis, the improvement wherein said magnetic circuit
~6~3Z~
comprises laminations layered in a direction perpendicular to
said axis whereby said magnetic circuit is substantially
incompressible in the direction of said axis.
The invention itself, however, together with additional
objects, features and advantages thereof, will be best understood
from the following description when read in conjunction with the
accompanying drawings in which:
Fig. 1 is a perspective view of one embodiment of the
linear-motor compressor hermetically encased within a protective
housing in accordance with the present invention;
Fig. 2 is an exploded perspective view of the linear
compressor motor and gas pump shown in Fig. l;
Fig. 3 is an elevational view of the compressor shown
in Fig. 1 taken partly in axial section along the line 3-3 of
Fig. l;
Fig. 4 is an axial sectional view of the compressor
shown in Fig. 1 in a stage of partial assembly;
Fig. 5 is a fragmentary axial sectional view of the
compressor shown in Fig. 3 in a second stage of partial assembly;
Fig. 6 is a graph used to explain the operation of the
compressor shown in Figs. 1 to 5;
Fig. 7 is a graph of the reluctance curve of one
embodiment of the compressor of Fig. l;
Fig. 8 is a fragmentary axial sectional view of an
alternative embodiment of the compressor of Fig. 1 which includes
a pair of zero-pitch internesting sprlngs;
Fig. 9 is an end view of the compressor of Fig. 8 which
-3a-
1~6920
shows the internested relation~hip of the springs in greater
detail
FIG. 10 i8 a plan view of a modified magnetic core which
may be used in the compressor of FI~. l;
FIG. 11 i8 a perspective view of the air-gap in the core
of FIG. 10 at an intermediate stage of fabrication;
FIG. 12 iS a perspective view of the completed air-gap
in the core of FIG. 10; and
FIG. 13 is a graph used to explain the operation of the
modified core shown in FIG. 10.
In the various figures, identical reference numerals
indicate identical parts. Referring to FIGS. 1-7, there is
qhown an exemplary embodiment of a linear-motor compressor
20 constructed pursuant to the disclosure of the aforementioned
U.S. Patent ~o. 3,947,155 which is disclosed herein to provide
a better understanding of how to make and use the features of
the present invention, which are disclosed in detail subse-
quently herein in conjunction with FIGS. 8-13, and represent
modifications to compressor 20. Compressor 20 is suspended
within a protective enclosure 21 by the suspension springs 22
which ideally provide a zero retarding force to the axial oscil-
latory movement of compres~or 20 and an infinite retarding force
to lateral or radial oscillatory movement thereof. Enclosure
21 i8 hermetically sealed and may be formed of sheet steel or
aluminum or molded plastic, and may assume a shape most con-
venient for the particular application. Lubricating oil is
preferably provided in a sump 23 at the bottom of the enclosure
at a depth qufficient to contact the lower portion of compressor
20. The oil will be splashed onto the moving part~ by the axial
reciprocating action of the tor. Alternatively, the oil may
be channeled to lubricated surfaces by other means known in the
art.
--4--
69ZO
The detailed description of compressor 20 may be hest
understood with reference to FIGS. 2-3. Magnetic circuit 30,
which includes magnetic core 31 and winding~ 32 and 33, ha~
a pair of spaced-apart poles 32a and 32b defining an air-gap
34 with the opposed surface~ or pole faces of pole~ 32a and
32b defining a portion of a frustoconical surface of revolu-
tion. Attached to opposite sides of the magnetic circuit by
means of bolts 35 and 36 is an outboard bearing and spring
retainer plate 37 and a cylinder block 38 having a pump chamber
ox cylinder 39 formed therein. As shown in FIG. 3, the taper
of air-gap 34 converges iD the direction of chamber 39 with the
center axis of gap 34 being coaxial with chamber 39.
Mbvable in air-gap 34 is a frustoconical armature 40 car-
ried by an armature rod 41. Armature 40 may be made of either
solid magnetically permeable material or stacked laminations
as shown. It has been found that the use of ~tacked laminations
increases the efficiency of the compressor by 15 per cent when
compared to an identical compressor with a solid magnetic core.
Armature rod 41 may be made of nonmagnetic material, such as
stainless steel, or, preferably, magnetically permeable material.
Mbunted on one end of rod 41 and slidable in chamber 39 is
a piston 42. For maximum compression efficiency, the sliding
clearance between piston 42 and the side wall of chamber 39
must be 8mRll: a nominal clearance of .0003 inches is preferred.
Mbunted in plate 37 is a sleeve bearing 43 disposed about
rod 41 at the end thereof remote from piston 42. Because of the
close ~liding clearance between piston 42 and the wall of
chamber 39, the piston will cooperate with bearing 43 to main-
tain rod 41 and armature 40 centered in air-gap 34 during axial
displacement of the armature, rod and piston.
Slidably mounted on plate 37 and clamping one end of a pair
of return springs 44 and 45 is an adjustable clamp bracket 46.
Bracket 46 may be tightly clamped to plate 37 by means of screw
_5_
c
6920
47 which iq threadably received in a split or slotted offset
portion of the clamp. The respective straight end-tang termina-
tions 44a and 45a at the outboard end of springs 44 and 45 are
clamped into as~ociated hole~ 46a and 46b of bracket 46 by means
of screws 48 and 49 which traverse associated bracket splits
leading to each of the clamp holes. One end of each of return
springs 44 and 45 is thus fixedly clamped in relation to magnetic
circuit 30 and air-gap 34. The other straight end-tangs 44b
and 45b of each return spring 44 and 45 is operatively clamped
to armature 40 by means of a spring clamp plate 50 which i8
mounted on rod 41 against the large diameter face of conical
armature 40.
End-tangs 44b and 45b are inserted into associated holes
50a and 50b respectively and clamped therein by means of screws
50c and 50d threadably received into respective split portions
of clamp SO. It should be ~oted that springs 44 and 45 are
coiled in the same direction but that each spring enters bracket
46 and clamp 50 from a direction 180 from the direction of
entry of the other. In this configuration the bending forces
imparted upon armature 40 by the spring~ during axial recipro-
cation of the armature tend to cancel each other, thereby assist-
ing bearing 43 and the bearing action of piston 42 to center
armature 40 in air-gap 34 during reciprocating axial movement.
It should be further noted that end-tangs 44a, 44b and 45a,
45b extend in a direction parallel to the central axis of the
spring~ from the periphery of the respective springs. This
feature allows bracket 46 and clamp 50 to assume a reduced
diameter, thereby reducing the required dimensions of enclosure
21 While several methods of terminating and affixing springs
44 and 45 will be evident to those skilled in the art, it ha~
been found that the use of straight end-tang terminations 44a,
44b and 45a, 45b on the return springs and the associated split
1~6~ZO
clamp mounting facilitate~ adjustment and assembly.
Compressor 20 is suspended in sump 23 a3 detailed above
with reference to FIG. 1. Because the lateral dimension of
block 38, that i~, the dimension perpendicular to a line between
bolts 35 and 36 and perpendicular to the axis of reciprocation,
is less than the corresponding lateral dimension of core 31,
which relationship is best seen in FIGS. l and 2, and because
of the access via side openings provided by the axial spacing
of block 38 from core 30 due to mounting pads 38c and 38d ~FIG.
2), oil splashed upwardly by the reciprocating action of com-
pressor 20 will enter the chamber 31a (FIG. 3) between core
30 and block 38. Splashed oil which contacts the minor diameter
face of armature 40 will be thrust into chamber 39 against the
back of piston 42 by the reciprocating action of the armature.
m is oil will lubricate the sides of chamber 39 in the area
of sliding contact with piston 42.
A valve plate and cylinder head assembly 51 is mounted on
cylinder block 38 by means of bolts 52. The suction and dis-
charge valves, the valve plate as~embly, and the cylinder head
may each be any one of the several ~tandard designs known to
the art and do not form a part of this invention. In a 450
BTU/Hr working embodiment of compressor 20 to be discussed in
detail hereinafter, valve plate assembly 51 is an adaptation
of the valve system from a commercially available Mbdel AE
Compressor manufactured by Tecumseh Products Company of
Tecumseh, Michigan. Valve assembly 51 will not be discussed
further except by reference during the discussion of the assembly
and operation of the compressor.
The economical method of assembling compressor 20 may be
best understood by reference to FIG. 4 in which compressor 20
i~ shown being assembled on an assembly surface 80. Cylinder
block 38 is fir~t placed head-end down on the assembly surface.
6~320
Then magnetic circuit 30 is loosely placed on the accurately
m~chined seating ~urfaces 38a and 3~3b of block 38 with the
respective bolt hole~ of the core and block roughly aligned.
The armature rod assembly, consisting of clamp S0, armature 40
and piston 42 all mounted on armature rod 41, is then seated
in the magnetic circuit by being piloted piston-end first into
chamber 39 until the piston extends sufficiently into chamber
39 such that the conical armature is seated against the pole
faces 32c and 32d which define conical air-gap 34. ~ote in
FIG. 4 that in this fully inserted condition piston 42 extends
beyond the head-end face 38c of cylinder block 38 by an amount
of distance indicated "b" when armature 40 abuts the pole faces.
The purpose of this extension will be explained in the discussion
of the operation of the compressor motor hereinafter. As the
armature is being thus seated, the geometry of the armature
and air-gap and the tight tolerance between the piston and
chamber wall causes the armature rod assembly to act as a set-
up jig which cams core 31 sideways so a~ to shift it laterally
on faces 38a and 38b to thereby automatically center the mag-
netic circuit and cylinder block with one another and with thearmature, rod and piston. The outboard bearing plate 37 and
bearing 43 is next mounted on the magnetic circuit, and then
bolts 35 and 36 are inserted through plate 37 and core 31 and
threaded into block 38, thereby automatically aligning bearing
43 with the co~n axis of the air-gap and compression chamber
and bringing the parts into accurated angular registry. Bolt~
35 and 36 may be then tightened down to secure the sub-assembly.
In the next stage of assembly shown in FIG. 5, end-tanys
44a, 44b and 45a, 45b of return springs 44 and 45 are inserted
and tightly clamped in adjustable bracket 46 and clamp 50.
Bracket 46 at this stage is loosely received on a mounting post
37a of plate 37 so that it can ve thereon as piston 42 i~
~96~ZO
raised to rest upon a jig block 81 which is inserted below the
piston in the pocket of the assembly surface 80. The piston
and armature will then be in the desired rest position, and
clamp 46 is then tightly clamped to post 37a after the valve
plate and cylinder head assemblies 51 are mounted to the cylinder
block, ~he motor will be ready for operation.
To operate the linear compressor motor, windings 32 and
33 must be connected to a source of alternating current. In
the embodiment of the invention illustrated herein, the sourc0
of alternating current is half-wave rectified utility power at
a frequency of 60 Hz. The motor thus operates at 3600 recipro-
cations per minute, It is well known in the art that maximum
~ compressor efficiency will be achieved when the resonant fre-
quency of the compressor during normal operation approaches the
line frequency of the exciting voltage. Thus, the natural
oscillating frequency of the piston, armature, rod and return
springs taken together with the normal suction and discharge
pressures in the compression chamber should approach 60 Hz.
me natural frequency of the return springs together with the
rod, piston and armature must, therefore, be less than the
frequency of exciting current. In the disclosed embodiment,
the natural frequency of the return springs and the rod, piston
and armature is preferably substantially equal to 38 Hz.
Operation of compres~or 20 may be best understood with
reference to FIG. 6 which is a timing diagram depicting the
relationships of selected parameters of compressor 20 during
one cycle of line voltage. The line voltage 60 describes a
substantiall~ sinusoidal pattern over the duration of a 360
cycle time. Because compressor 20 presents an inductive load
to line voltage 60, it is to be expected that the current 61
will lag voltage 6~ and describe a rectified half wave which is
periodic but not sinusoidal. The flux 62 through magnetic circuit
z~
30 follows, but slightly lags, current 61. me ordinates of
voltage 60, current 61 and flux 62 are measured in units of
volts, amps and kilomaxwells respectively and are not to scale.
However, voltage 60, current 61 and flux 62 have a common zero
ordinate reference for clarity of understanding. The armature
displacement 63 is measured-in units of inches with the zero
displacement reference being the abutment position of armature
40 against pole faces 32c and 32d which reference position is
depicted in FIG. 4. The magnetic force 64 is measu~ed in units
of pounds with reference to positive displacement of armature
40. Thus, magnetic force 64 which tends to move armature 40 in
a negative direction, that is, a direction toward the zero dis-
placement reference, i8 shown executing a negative excursion
from the zero magnetic reference point. Similarly, spring
force 65, which is the force exerted upon armature 40 by springs
44 and 45, and pressure force 66, which is the force exerted
on the compression face of piston 42, are measured in units of
pounds with reference to a positive axial displacement armature
40, that i8, a spring or pressure force which tends to move
armature 40 in the direction of positive axial displacement is
considered to be a positive force. Magnetic force 64, spring
force 65 and pressure force 66 have a common zero ordinate
reference for clarity of understanding. The abscissa of FIG. 6
is measured in units of electrical time in degrees of a single
cycle of line voltage 60. It should be noted with respect to
FIG. 6 that, while the signals shown therein are not to scale,
the geometry of each signal is duplicated from test results
based upon the 450 BTU/Hr working embodiment to be set forth in
detail hereinafter. -
In the operation of compressor 20 voltage 60 begins a
positive excursion at electrical time zero degrees and induces ~-
current 61 in the windings of magnetic circuit 30~ Current 61
--10--
~6~2~
induces, in turn, flux 62 in core 31 and armature 40. Thu~,
starting at zero degree~ electrical time, magnetic force 64
gradually increases (in the negative direction) and urges armature
40, and therefore pi~ton 42, in the negative displacement direc-
tion. It will be noted from ~IG. 6 that, at time zero degrees,
armature 40 is moving in the positive displacement direction
which means that, at the beginning of an electrical cycle, the
armature is executing its return stroke, as opposed to it~
compres~ion stroke, as a result of the momentum imparted to
the moving assembly comprising armature 40, rod 41, piston 42
and clamp 50 by return springs 44 and 45 during the preceding
electrical cycle. Spring force 65 is negative at time zero
degrees indicating that springs 44 and 45 are in compression
and exert a force on armature 40 in the negative displacement
direction. Thu~, shortly after time zero degrees, magnetic
force 64 cooperates with spring force 65 to work against the
momentum of the assembly to arrest positive displacement thereof
and begin movement in the negative direction.
At an electrical time of 90 degrees, displacement 63 has
reached its maximum value and the moving assembly has reached
its "top dead point" of operation. The assembly will begin
to move in the negative direction. As is to be expected, at
time 90 degrees ~pring force 65 has reached its maximum negative
or compression value and will begin to move in the positive
direction. Magnetic force 64 will continue to increase in a
negative direction as current 61 and resulting flux 62 increases.
Armature 40 and piston 42 now move in the negative displacement
or working direction toward the head-end of pump chamber 39,
compressing the ga~ in chamber 39 to a desired discharge pressure
at which the discharge valve will open.
~ hen moving in the negative displacement direction, armature
40 will eventually pass its neutral position so that spring3 44
('' !
and 45 go into tension and begin to retard further negative
displacement of the moving assembly. In FIG. 6 this neutral
or zero spring force position i~ achieved at an electrical
time of approximately 208 degrees. It should be noted that
at time 208 degrees flux 62 has already passed its maximum point
and has begun to decline toward zero.
When magnetic force 64 and the rate of change of momentum
of the moving mass 40, 41, 42 and 50 i~ equal to the sum of
spring force 65 exerted on armature 40 by return springs 44 and
45 in tension and pressure force 66 exerted on the face of
piston 42 by the compressed gas in chamber 39, positive dis-
placement is arrested and the armature and piston reach their
"bottom dead point" of operation. In FIG. 6 this occurs at an
electrical time of approximately 265 degrees. It should be
noted that at this "bottom dead point" time flux 62 in magnetic
circuit 30 i8 less than half of its maximum value.
Magnetic force 64 will continue to decline after bottom
dead point time 265 degrees so that spring force 65 and pressure
force 66 govern movement of the armature and piston and return
the moving assembly in the positive displacement or return
direction. Winding current 61 reaches a zero value at time
320 degrees. Because the current is rectified, voltage 60
returns to zero at this time. The moving assembly comprising
armature 40, rod 41, piston 42 and clamp 50 continue motion
in the positive displacement or return direction under the
influence of pre~sure force 66 and spring force 65. Positive
displacement 63 will continue to increase until the moving ~--
asse~bly reaches its top dead point of operation under the
influence of the 3pring and magnetic forces as outlined above.
In prior linear-motor compressors of the type which include
a cylindrical armature and air-gap, the armature is attracted
into the air-gap and made to do work until it reaches a point
-12-
1~6~
at which its top and bottom end faces are flush with the faces
defining the axially opposite ends of the air-gap. At this
point, the armature completely fills the air-gap and, since
the air space between the armature and pole faces i5 constant,
the reluctance of the total magnetic circuit is at a minimum.
The armature can thus be made to do no further work in that
cycle. It has apparently been assumed by others in the art
that this constraint will also apply to a linear motor having
a conical armature and air-gap; this, however, is not the case.
Indeed, maximum compressor efficien y is obtained when the
conical armature "fills the air-gap' at the point of maximum
flux and, since this maximum flux point will not necessarily
occur at the "bottom dead point~ of operation, it is advan-
tageous to have the armature continue through the air-gap beyond
this flush point. Since the air space between the conical
armature and opposing pole faces is no longer constant and is,
in fact, a function of axial displacement, the reluctance of the
total magnetic circuit will continue to decrease even though
part of the armature is moving out of the air-gap.
Returning to FIG! 4, it can be seen that armature 40 extends
out of the air-gap a distance "a" when piston 42 extends a
distance "b" beyond the end face of cylinder block 38. Ihe flush
condition will exist when the minimum diameters of the armature
and air-gap are coplanar--i.e., when a ~ 0. It is undesirable
to allow armature 40 to strike the pole faces; for this reason,
distance "a" is made much larger than distance "b". The piston
will thus strike the valve plate before the armature can reach
the pole faces, which prevents the armature from striking the
pola faces.
Referring to FIG. 6, it will be Qeen that the "bottom
dead point" of operation is achieved at an electrical time of
-13-
1~69ZO
about 265 degree~. At this time flux 62 in magnetic circuit 30
i~ le~s than half of its maximum value. Armature 40 i~ to be
positioned on rod 41 so that the armature i8 flush with pole
piece~ 32a and 32b at an electrical time of approximately 180,
at which time flux 62 achieves its maximum value. This may be
accomplished by modifying the diameter of air-gap 34 vis-a-vis
the diameter of armature 40, while maintaining identical included
angles of taper, so that, when piston 42 is in the set-up position
shown in FIG. 4, armature 40 extends through the air-gap a
distance calculated to achieve the desired flush position at the
desired time based upon the test results ~hown in FIG. 6. Re-
ferring again to FIG. 4, in the 450 BTU/~r working embodiment
of the compressor, armature 40 is positioned to extend approxi-
mately .350 inches beyond pole pieces when piston 42 extends
.030 inche~ beyond the head-end of cylinder block 38.
me minimum air space between the pole faces and the armature
will exist when the piston abuts the valve plate assembly. In
the disclosed embodiment this minimum space, that is, the minimum
distance from a pole face to the armature as measured in a direction
perpendicular to the pole face, is substantially .0035 inches.
It would, of course, be undesirable to allow the piston to
continually strike the valve plate during normal operation.
However, as is well known in the art, compression efficiency
i9 optimized when the distance between the piston face and the
valve plate approaches zero at the "bottom dead point" o,f opera-
tion. Magnetic force, spring force and compression force must
be thus optimized to achieve maximum compression efficiency with-
out allowing the pîston to strike the valve plate.
While it has been stated for purposes of explaining the
operation of the compressor of FIGS. 1-5 that the armature
moves "into" and "out of" the air-gap, it will be appreciated
from the discussion immediately above that the armature need
-14-
not move "entirely out of the air-gap" nor for that matter
need "a major portion thereof" be located outside of the air-gap
at the "top dead point" of operation, contrary to the dis-
closur~s in the United States Barthalon patents 3,542,495
and 3,461,806 respectively. Indeed, in the embodiment dis-
closed herein, which operates at 450 BTU/Hr at standard rating
point conditions, the total compression stroke is only .8
inches, and the armature exposure at the "top dead point" of
operation is less than 50 per cent
When the magnetic circuit reluctance characteri~tics de-
tailed above have been defined--i.e., a substantially linear
reluctance curve over the entire stroke length and an armature
flush condition at the time of maximum flux--then the included
angle of taper of armature 40 and air-gap 34 may be specified.
As stated above, it has been found that, under the above re-
cited conditions, a piston extension dimension "b" of .030
inches yields good results. To achieve this dimension, the
included angle of taper of the armature and air-gap should be
at least 10, and a range of taper included angles between 10
and 14 i8 preferred.
The aforementioned Barthalon patents teach that the
efficiency of a linear motor will be optimized if the reluc-
tance of the magnetic circuit varies linearly with armature
movement As the aforementioned U. S. Patent No. 3,947,155
teaches, the stability of a pump which may occasionally
operate below atmospheric pressure, such as a refrigeration
compressor, will be enhanced if the linear reluctance curve
also has a low slope. The various design parameters have been
optimized in the present compressor motor to achieve this
desired result. While it is not necessary to have the angle
~l~g6~0
of taper of the armature identical to that of the air-gap, it
has been found that this condition gives the best overall
results. It ha~ also been found that the best results are
achieved if the net cross section of the armature, that is, the
cross sectional area of the armature taken on a plane through
the center of the armature parallel to the axis of movement
and excluding the armature rod, is equal to about 80 per cent
of the effective cross sectional area of the pole piece. The
effective cross sectional area of the pole piece is that area
taken on a plane parallel to the axis of movement of the arma-
ture and perpendicular to the flux through the pole piece.
The shape of this cross section should be substantially ~quare
rather than rectangular to achieve the minimum winding length
per unit of desired flux. The gross cross sectional area of the
armature, tXat is, the cross sectional area of the armature
taken as above but including the armature rod, should be
greater than the effective cross sectional area of the pole
piece. This arrangement yields good re3ults, particularly when
an armature rod of magnetically permeable material is used to
increase the "magnetic cross section" of the armature.
The reiuctance curve of the above-mentioned 450 BTU/Hr
embodiment is shown in FIG. 7. In the curve 70 of FIG. 7 the
abscissa is in incheY of displacement as measured from the
condition of FIG. 4 when the armature is seated in the magnet-
ic core. The ordinate measurement of reluctance indicates
that minimum reluctance at the position of FIG. 4 is approxi-
- mately .001 ampere-turns per maxwell. It has been found that
an excessive slope angle 71 is accompanied by frequent impact
of piston 42 upon valve plate 51, while an insufficient slope
results in loss of mechanical efficiency and a reduced range
of conditions for successful operation. It will be noted
-16-
~69Z~)
that reluctance curve 70 is substantially linear over the
entire stroke of .8 inches and has a slope of approximately
.022 ampere-turns per maxwell-inch. ~he parameters of this
450 BTU/Hr working embodiment which contribute to this low-
sloped, linear reluctance curve, and the consequent high com-
pressor efficiency, are set forth in the discussion of the
working embodiment detailed hereinafter.
An alternative to the three-turn paired spring arrangement
in the compressor of FIG. 1 is shown in FIGS. 8 and 9. An
outboard bearing and spring retainer plate 100 i3 clamped to
magnetic circuit 30 and cylinder block 38 by the tie bolts
102 and 104. Plate 100 has a pair of spring retainers 106
and 108 each of which fixedly clamps one end of the zero pitch
linear springs 110 and 112. Respective straight end-tang ter-
mination~ llOa and 112a at the outboard end of springs 110 and
112 are clamped into associated holes 106a and 108a of clamps
106 and 108 by means of screws 114 and 116 which traverse
associated bracket splits leading to each of the clamp holes.
End-tangs llOb and 112b are similarly clamped to armature 40
by means of spring clamp plate 50.
It will be appreciated by those skilled in the art that,
depending upon the manufacturing technique used to fabricate
the spring~, a "zero pitch" spring will have a pitch between
zero and the diameter of the spring material. Where straight
end-tangs are required, the spring is usually first coiled
on a circular mandrel or jig with the end-tangs extending tan-
gentially from the coil. The end-tangs are then bent to
posi ions perpendicular to the plane of the coil. The pitch
of the spring thus formed will be substantially equal to zero
within some tolerance range which depends upon the resilience
10~6920
of the material used to wind the spring.
There are approximately .92 turns of spring material in
springs 110 and 112. End-tangs llOa and llOb of ~pring 110 are
thus laterally spaced from each other allowing room for spring
112 to pass therethrough before terminating in clamp 50. Sim-
ilarly, end-tang~ 112a and 112b are spaced to allow pa~sage
of spring 110 therebetween, thereby internesting the springs.
In this geometry the coils of springs 110 and 112 are aliyned
with a line connecting $ie bolts 102 and 104 rather than being
perpendicular therewith and are contained within the lateral
perimeter of compressor 20 defined by mag~etic circuit 30,
thereby reducing the lateral and axial dimensions of the com-
pressor. Furthermore, with the coils of springs 110 and 112
disposed in axial proximity to magnetic circuit 30, housing 21
which encompas~es compressor 20 may assume an eliptical shape
which is believed to reduce the level of acoustical noise
emanating from an operating unit.
The zero pitch internesting springs shown in FIGS. 8 and
9 have the additional advantage of reducing the twisting forces
imparted upon armature 40 almost to zero. This reduction in
the torsion or twisting forces on the armature and springs
results in long spring life and helps maintain armature 40
within air-gap 34 during axial reciprocation thereof.
As shown in FIG. 2, magnetic core 31 comprises stacked
laminations attached in a manner well known in the art. Al-
ternatively and preferably, the magnetic core may be comprised
of first and second inner loops spirally wound of magnetic
strip material with the loops placed in abutment and banded
together by an outer loop of the same magnetic strip material.
Such a core 120 is shown in FIG. 10 and is constructed by
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first separately winding a pair of identical inner loops 122
and 124 of magnetic strip material to form spiral wrap
pattern 126. When loop 122 has reached the desired thickness,
the strip material may be terminated and tacked as shown at
128. When loop 124 has reached the desired thickness, the
~trip material is to be tacked as at 130, but need not be ter
minated. Loops 122 and 124 are then placed in flat end abut-
ment on plane 129 and the magnetic strip material extending
from tack 130, or a separate strip material tacked onto either
loop at a convenient attachment point, is wound around the ex-
posed periphery of the dual loop subassembly to form an outer
convoluted loop 132 which holds inner loops 122 and 124 tightly
together as disclosed in U.S. Patent 2,431,128, which issued
on November 18, 1947 in the name of E. A. Link. Conical air-
gap 134 is then machined in the area of abutment of inner loops
122 and 124. Windings 32 and 33 will be wound about the opposing
pole pieces and will have magnetic communication carried entirely
by the inner loops. For this reason, outer loop 132 may be of
any convenient material. The magnetic core shown in FIG. 10
is more easily assembled and has less waste material than stacked
lamination ~ore 31.
FIG. 11 is a perspective view of conical air-gap 134 after
the air-gap is first machined into the area of abutment of
first and second loops 122 and 124. When the minor diameter
of gap 134 is le~s than the width of the core (i.e., the dimen-
sion perpendicular to plane 129), then the pole pieces 136 and
138, rather than being isolated from each other, are connected
by the magnetic bridges or connection~ 140 and 142 on either
side of the machined gap. In order to prevent a short in the
magnetic circuit and to facilitate the mounting of windings
--19--
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32 and 33 upon pole piece-~ 136 and 138, connections 140 and
142 must be removed at a second machining stage in the fabri-
cation of magnetic core 120.
FIG. 12 is a perspective view of air-gap 134 in magnetic
core 120 after bridges 140 and 142 have been removed. Bridges
140 and 142 have been removed by machining acro~s the faces of
pole pieces 136 and 138 in a pair of planes X and Y respec-
tively perpendicular to the central axi~ of poles 136 and 138
and parallel to but displaced on opposite sides of the axi~ ;
of reciprocation. When the distance between planes X and Y
is less than the maximum diameter of gap 134, this machining
will produce in these planes the triangular coplanar exposed
gap surfaces 140x, 142x, and 140y, 142y upon oppo~ing faces of
pole piece~ 136 and 138 respectively. When the gap between
planes X and Y is to be only sufficient to allow insertion of
windings 32 and 33, a distance between the planes of 16.7 per
cent of the cross-sectional area of the poles is sufficient,
or, in other words, the ratio of the distance between the
planes and the cross-sectional area of the poles is 0.167 ins. 1.
~owever, it has been discovered pur~uant to the present
invention that compressor operation is enhanced when the dis-
tance between planes X and Y is increased beyond this 16.7 per
cent figure. In a specific 450 BTU/Hr working embodiment of
the present invention having 1.5 inch-square poles, the dis-
tance between planes X and Y was increased to .8 inches or
approximately 35.5 per cent of the cross-sectional area of
the poles. This arrangement yielded the result~ shown in FIG.
13 when compared to a similar 450 BTU/Hr unit with a planar
gap of .375 inches or 16.7 per cent. In FIG. 13 BTU/Hr output
i8 plotted versu~ evaporation temperatura. Da~hed curve 150
depicts the output of the .375 inch unit over a wide range of
evaporation temperatures while curve 152 represents the output
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~ ~Q~i9~)
of the .8 inch unit over the same range~ It can be seen that
the two units perform equally at rating point conditions--point
154--and perform similarly at evaporation temperatures lower
than rating point. However, at higher evaporation temperatures
the performance of the .375 inch unit falls off much more rap~dly
than the performance of the .8 inch unit. It should be noted
that the curves of FIG. 13 were plotted from actual test results
and are to scale.
Strip wound core 120 may replace laminated core 31 in
compressor 20 of FIG. 2. In this preferred compressor assembly,
tie bolts 35 and 36 pass through a pair of substantially tri-
angular apertures 131 and 133 which are formed in the area of
abutment of inner loops 122 and 124 and are bounded by the
inner loops and outer loop 132 as be~t seen in FIG. 10. Aper-
tures 131 and 133 afford core 120 a greater degree of lateral
"slop" in the assembly stage, thus facilitating the automatic
alignment process discussed above with respect to FIG. 4. In
addition, since the individual wraps or laminations of core 120
are layered in a direction perpendicular to the axis of recip-
rocation and to the axes of tie bolts 35 and 36, the retention
stres~es exerted by the tie bolts are taken edgewise by each
wrap instead of parallel to the thickness dimension of each
wrap. Hence, the core laminations or wraps are not subjected
to bolt forces tending to squeeze them together. Due to thi~
orientation of the bolts parallel to the lateral or width
dimention of the wraps, the strip wound core is not compressible
in the direction of tightening of the tie bolts. For this
reason, it is easier to hold alignment tolerances when core 120
is used.
The material disclo~ed above with reference to FIGs. 1 to
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1(9~6920
7, as well as the general concept of fabricating the air-gap
as shown in FIGS. 11 and 12, are the subject of the afore-
mentioned U. S. Patent No. 3,947~155. ~his material is di~-
closed herein to facilitate understanding of the present
invention and because it is the best method presently known r
for practicing the invention.
Several working embodiments of compressor 20 have been
built and tested, one such embodiment i5 the 450 BTU/Hr (nominal)
unit mentioned above and drawn to scale in FIGS. 1-5. By way
10of example and not by way of limitation, the parameters which
contribute to the 1OW-Q1OPe linear reluctance curve and the
resulting high compressor efficiency at rating point conditions
are as follows:
mass of piston 42. . . . . . . 0.17 lbm
mass of armature 40. . . . . . 0.8 lbm
mass of rod 41 . . . . . . . . 0.13 lbm
mass of clamp 50 . . . . . . . 0.12 lbm
effective mass of springs 44
and 45 (1/3 actual mas8) . . 0.08 lbm
rate of ~prings 44 and 45. . . 200 lb/in
material of rod 41 . . . . . . 1060 steel
net cross-~ectional area of
armature 40. . . . . . . . . 1.76 sq. in.
gross cross-sectional area
of armature 40 (and 41). . . 2.32 sq. in.
effective cross-sectional
area of pole pieces 32a and
32b. . . . . . . . . . . . . 2.25 sq. in.
resistence of windings 32
and 33 . . . . . . . . . . . 2.10 ohms
number of turns in windings
32 and 33. . . . . . . . . . 400
refrigerant suction pressure . 4.4 psig
refrigerant discharge pressure 180 psig
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. ,
,
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1(9~69120
refrigerant temperature enter-
ing compre~sor housing . . . 90F
bore . . . . . . . . . . . . . 1.156 inches dia.
flux path area . . . . . . . . 2.25 sq. in.
In the working embodiment with the above exemplary parameters,
the following results were measured at refrigeration industry
standard rating p~int conditions after 10,000 hours of oper-
ation:
capacity . . . . . . 485 BTU/Hr
power input. . . . . 134 watts
efficiency . . . . . 3.62 BTU/watt-hour (Weston)
In addition, the following results, which are difficult to
accurately measure in a working linear compressor, were cal-
culated rom a computer analysis of the 450 BTU/Hr model, the
analysis being similar to that set forth above with reference
to FIG. 6:
length of stroke . . . . . . . 0.54 in.
position of A/C power
cycle at "top dead
point~ of operation. . . . . 91 degrees
position of A/C power
cycle at flush position. . . 207 degrees
current at flush position. . . 4.9 amps
flux at flush position . . . . 213 kilomaxwells
position of A/C power
at maximum flux. . . . . . . 180 degree~
current at maximum flux. . . . 7 amps
maximum flux . . . . . . . . . 231 kilomaxwells
spring force at "top
dead point" of
operation. . . . . . . . . . -70 lbf
spring force at "bottom
dead point" of operation . . +38 lbf
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~6g2~
position of A/C power at
opening of discharge
valve. . . . . . . . . . . . 252 degrees
AR discussed above, reluctance curve 70 at FIG,7 indicates
that this embodiment achieved the objective of having a low-
sloped, linear reluctance curve. Furthermore, the above data
indicates that the objective of achieving maximum flux at
the flush position has been achieved within 8 per cent.
From the foregoing description, it will now be apparent
that there has been provided, in accordance with the invention,
an improved moving-iron linear compressor motor that fully
satisfies the objects and advantages set forth above. While
the invention has been described in conjunction with specific
embodiments thereof, it is evident that many alternatives,
modifications, and variations will be apparent to those skilled
in the art in light of the foregoing description. It will be
further apparent that, while the invention has been disclosed
and exemplified in connection with a refrigeration system,
the invention is equally applicable to other types of refrig-
erant systems and that, indeed, many principles of the inven-
tion may be applied generally to gas pumps, such as air com-
pressors or the like. Accordingly, the invention is intended
to embrace all such alternatives, modifications, and varia-
tions as fall within the spirit and broad scope of the
appended claims.
This application is a divisional application of
copending Canadian Patent Application Serial No. 332,895
filed July 31, 1979.
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