Note: Descriptions are shown in the official language in which they were submitted.
~05~733
This invention relates to reciprocating
electric motors of the type in which a magnetically
permeable armature is disposed to axially recipro-
cate within the air-gap of a fixed electromagnetic
S circuit.
In the art of moving-iron linear-motor
compressors, much effort has been expended without
having achieved significant commercial success (see
P. W. Curwen, "Recent Developments of Oil-Free Linear-
~otor Resonant-Piston Compressors," ~SME publication
69-FE-36, June, 1969). The linear-motor compressor
disclosed herein has been subjected to extensive lab-
oratory testing and the design parameters have been
verified through the use of iterative computer pro-
gramming techniques, and therefore, the requirementsof a commercially viable product are believed to have
at last been achieved.
The United States sarthalon Patent 3,461,806
teaches that the efficiency of a linear motor will be
optimized if the reluctance of the magnetic circuit
varies linearly with armature movement. Pursuant to
the present invention, it has been discovered that,
in a pump which may act below atmospheric pressure,
such as a refriyeration compressor, stability will be
enhanced if the reluctance curve has a low slope. It
is, therefore, an object of the present invention to
provide a moving-iron linear-motor compressor having
not only a substantially linear reluctance curve but
also one of low slope.
1. ~
~53733
It is another object of the present inven-
tion to provide a moving-iron linear-motor compressor
that is easy and economical to assemble.
It is a further object of the present in-
vention to provide a moving-iron linear-motor com-
pressor in which radial deflection of the armature
rod is reduced or eliminated.
It is yet a further object of the present
invention to provide a moving-iron linear-mo~,or com-
pressor in which axial movement of the piston withrespect to the armature during operation is prevented.
To achieve a linear reluctance curve of
low slope, the present invention provides a moving-
iron linear-motor compressor in which the armature
and the air-gap defined by the pole pieces of the
core of the magnetic circuit have a conical geometry,
and preferably the same axial dimension, However,
the minimum diameter of the air-gap is greate,r than
the minimum diameter of the armature so that the
armature may move through the air-gap a substantial
distance beyond the point where these minimum diameters
are coplanar. This relationship is coordinated with
the electromagnetic drive such that the armature may
~e flush with the pole pieces at the time of maximum
flux through the magnetic circuit to thereby optimize
the performance and efficiency of the motor and pump.
In addition, the armature rod is made of magnetically
permeable material.
~OS3733
The compressor may be easily and quickly assembled
in a "layered" fashion, i.e., a sequential part stack-up
assembly procedure. An additional feature resulting from this
layered construction and from the conical armature geometry is
that insertion of the piston, rod and armature into the cylinder
block and magnetic circuit, such that the piston is wi~hin th~
compression chamber and the armature is seated against the pole
faces, automatically aligns the cylinder block and the magnetic
circuit with the piston, rod and armature.
To reduce armature rod deflection, the return .neans
are connected to the armature between the piston and an outboard
bearing. To further reduce deflection, the return means inclu-
des a pair of complementary return springs with straight end-
tangs.
Axial movement of the piston with respect to the
armature is prevented by providing a piston element having a
positive abutment interconnection with the armature.
Accordingly, the present invention provides in an
electromagnetic circuit of the type in which a magnetically per-
meable armature is disposed to axially reciprocate within anair-yap defined by a pair of spaced-apart, coaxial poles on a
fixed magnetic core, the improvement wherein said air-gap
comprises 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 planar 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 invention itself, however, together with
~ _3_
~05;~73;~
additional objects, features and advantages thereof, will be
best understood from the following description when read in
connection with the accompanying drawings in which:
- 3a -
lOS3733
FIG. 1 is a perspective view of one embodiment
of the linear-motor compressor hermetically encased with-
in a protective housing in accordance with the present
invention;
S 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 ~long the
line 3-3 of FIG. l;
FIG. 4 is an axial sectional view of the com-
pressor shown in FIG. 3 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 opera-
tion of the compressor shown in FIGS. 1-5;
FIG. 7 is a graph of the reluctance curve of
one embodiment of the compressor of FIG. l;
FIG. 8 is a seGtional view of an alternative
piston assembly which may be used in the compressori
FIG. 9 is a perspective view of an alterna-
tive paired spring arrangement which may be used in the
compressor;
FIG. 10 is a fragmentary axial sectional view
2S of an alternative embodiment of the compressor of FIG. 1
which includes a pair of zero-pitch internesting springs;
~C53733
FIG. 11 is an end view of the compressor
of FIG. 10 which shows the internested relationship of
the springs in greater detail;
FIG. 12 is a plan view of a modified magnetic
core which may be used in the compressor of FIG. 1,
FIG. 13 is a perspective view of the air-gap
in the core of FIG. 12 at an intermediate stage of
fabrication;
FIG. 14 is a perspective view of the completed
air-gap in the core of FIG. 12; and
FIG. 15 iS a graph used to explain the opera-
tion of the modified core shown in FIG. 12.
In the various figures, identical reference
numerals indicate identical parts. Referring to FIG.
1, there is shown an exemplary embodiment of a linear-
motor compressor 20 constructed pursuant to the present
invention suspended within a protective enclosure 21
by the suspension springs 22 which ideally provide a
zero retarding force to the axial oscillatory move-
ment of compressor 20 and an infinite retarding forceto lateral or radial oscillatory movement thereof.
Enclosure 21 is hermetically sealed and may be formed
of sheet steel or aluminum or molded plastic, and may
assume a shape most convenient for the particular
application. Lubricating oil is preferably provided
in a sump 23 at the bottom of the enclosure at a depth
sufficient to contact the lower portion of compressor
~053733
20. The oil will be splashed onto the moving parts
by the axial reciprocating action of the motor. Al-
ternatively, the oil may be channeled to lubricated
surfaces by other means known in the art.
The detailed description of compressor 20
may be best understood with reference to FIGS. 2-3.
Magnetic circuit 30, which includes magnetic core 31
and windings 32 and 33, has a pair of spaced-apart
poles 32a and 32b defining an air-gap 34 with the
opposed surfaces or pole faces of poles 32a and 32b
defining a portion of a frustoconical surface of
revolution. Attached to opposite sides of the mag-
netic circuit by means of bolts 35 and 36 is an out-
board bearing and spring retainer plate 37 and a
cylinder block 38 having a pump chamber or cylinder
39 formed therein. As shown in FIG. 3, the taper of
air-gap 34 converges in the direction of chamber 39
with the center axis of gap 34 being coaxial with
chamber 39.
Movable in air-gap 34 is a frustoconical
armature 40 carried by an armature rod 41. Armature
40 may be made of either solid magn~tically permeable
material or, preferably, stacked laminations as shown.
It has been found that the use of stacked laminations
increases the efficiency of the compressor by 15 per
cent. Armature rod 41 may be made of nonmagnetic
material, such as stainless steel, or, preferably,
magnetically permeable material.
~353733
Mounted on one end of rod 41 and slidable
in chamber 39 is a piston 42. For maximum compression
efficiency, the slidiny clearance between piston 42
and the side wall of chamber 39 must be small: a
nominal clearance of .0003 inches is preferred.
~ ounted in plate 37 is a sleeve bearing 43
disposed about rod 41 at the end thereof remote from
piston 42. Because of the close sliding clearance
between piston 42 and the wall of chamber 39, the
piston will cooperate with bearing 43 to maintain
rod 41 and armature 40 centered in air-gap 34 during
axial displacemellt of the armature, rod and pistorl.
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 47 which
is threadably received in a split or slotted offset
portion of the clamp. The respective straight end-
tang terminations 44a and 45a at the outboard end
of springs 44 and 45 are clamped into associated holes
46a and 46b of bracket 46 ~y 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 is mounted on
rod 41 against the large diameter face of conical
armature 40.
~053733
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 S0.
It should be noted 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 configu-
ration the hending forces imparted upon armature 40
by the springs during axial reciprocation of the
armature tend to cancel each other, thereby assisting
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 springs 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 terminatingand affixing springs 44 and 45 will be evident to
those skilled in the art, it has been found that the
use of straight end-tang terminations 44a, 44b and 45a,
45b on the return springs and the associated split
clamp mounting facilitates adjustment and assembly.
Compressor 20 is suspended in sump 23 as
detailed above with reference to FIG. 1. Because the
lateral dimension of block 38, that is, the dimension
perpendicular to a line between bolts 35 and 36 and
perpendicular to the axis of reciprocation, is less
~ S3733 ~
than the corresponding lateral dimension of core 31,
which relationship is best seen in FIGS. 1 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 compressor 20 will enter the
chamber 31a tFIG. 3) between core 30 and block 38.
Splashed oil which contacts the minor diameter face of
armature 40 will be thrust into chamber 39 ayainst the
back of piston 42 by the reciprocating action of the
armature. This 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 discharge valves, the valve plate assem-
bly, and the cylinder head may each be any one of the
several standard designs known to the art and do not
form a part of this invention. In a 450 BTU/Hr working
embodiment of the present invention to be discussed in
detail hereinafter, valve plate assembly 51 is an adapta-
tion of the valve system from a commercially ~vailable
Model AE Compressor manufactured by Tecumseh Products
Company of Tecumseh, ~ichigan. 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 the
compressor provided by the present invention may be
best understood by reference to FIG. 4 in which com-
pressor 20 is shown being assembled on an assembly
1~S3733
surface 80. Cylînder block 38 is first placed head-
end down on the assembly surface. Then magnetic cir-
cuit 30 is loosely placed on the accurately machined
seating surfaces 38a and 38b of block 38 with the
respective bolt holes of the core and block roughly
aligned. The armature rod assembly, consisting of
clamp 50, armature 40 and piston 42 all mounted on
armature rod 41, is then seated in the magnetic cir-
cuit 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. Note 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 cham-
ber wall causes the armature rod assembly to act as
a set-up jig which cams core 31 sideways so as to
shift it laterally on faces 38a and 38b to thereby
automatically center the magnetic circuit and cylin-
der block with one another and with the armature, rod
and piston. The outboard bearing plate 37 and bearing
10 .
~053733
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 automati-
cally aligning bearing 43 with the common axis of
the air-gap and compression chamber and bringing the
parts into accurated angular registry. Bolts 35 and
36 may be then tightened down to secure the sub-assem-
bly.
In the next stage of assembly shown in FIG.
5, end-tangs 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 move thereon as piston 42 is 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, the 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 inven-
tion illustrated herein, the source of alternatingcurrent is half-wave rectified utility power at a
frequency of 60 Hz. The motor thus operates at 3600
reciprocations per minute. It is well known in the
art that maximum compressor efficiency will be achieved
when the resonant frequency of the compressor during
normal operation approaches the line frequency of the
ios~733
exciting voltage. Thus, the natural oscillating fre-
quency 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. The natural frequency of the return springs
together with the rod, piston and armature must,
therefore, be less than the frequency of exciting
curren~. In the disclosed embodiment, the natural
frequency of the return springs and the rod, piston
and armature i5 preferably substantially equal to 38 Hz.
Operation of compressor 20 may be best
understood with reference to FIG. 6 which is a timing
diagram depicting the relationships of selected para-
meters of compressor 20 during one cycle of line vol-
tage. The line voltage 60 describes a substantiallysinusoidal pattern over the duration of a 360 cycle
time. Because compressor 20 presents an inductive
lead to line voltage 60, it is to be expected that
the current 61 will lag voltage 60 and describe a
rectified half wave which is periodic but not sinusoidal.
The flux 62 through magnetic circuit 30 follows, but
slightly lags, current 61. The 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
12.
~053733
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
measured 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 direc-
tion, that is, a direction toward the zero displace-
ment reference, is shown executing a negative excur-
sion from the zero magnetic reference point. Similar-
ly, spring force 65, which is the force exer~ed uponarmature 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 is, 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 there-
in 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.
-
~053733
In the operation of compressor 20 voltage 60
begins a positive excursion at electrical time zero de-
grees and induces current 61 in the windings of magnetic
circuit 30. Current 61 induces, in turn, flux 62 in
core 31 and armature 40. Thus, starting at zero degrees
electrical time, magnetic force 64 gradually increases
(in the negative direction) and urges armature 40, and
therefore piston 42, in the negative displacement direc-
tion. It will be noted from FIG. 6 that, at time zero
degrees, armature 40 is moving in the positi~e displace-
ment direction which means that, at the beginning of
an electrical cycle, the armature is executing its re-
turn stroke, as opposed to its compression 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 4~ and 45 during the preceding elec-
trical cycle. Spring force 65 is negative at time zero
degrees indicating that springs 44 and 45 are in com-
pression and exert a force on armature 40 in the nega-
tive displacement direction. Thus, 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 move-
ment in the negative direction.
At an electrical time of 90 degrees, dis-
placement 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
14.
-
~053733
negative direction. As is to be expected, at time
90 degrees spring force 65 has reached its maximum
negative or compression value and will begin to move
in the positive direction. ~agnetic force 64 will
continue to increase in a negative direction as cur-
rent 61 and resulting flux 62 increase. Armature 40
and piStOll 42 now move in the negative displacement
or working direction toward the head-end of pump cham-
ber 39, compressing the gas in chamber 39 to a de-
sired discharge pressure at which the discharge valve
will open.
When moving in the negative displacement
direction, armature 40 will eventually pass its
neutral position so that springs 44 and 45 go into
lS tension and hegin to retard further negative displace-
ment of the moving assembly. In FIG. 6 this neutral
or zero spring force position is achieved at an elec-
trical 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 is equal to the sum of spring force 65 exerted
on armature 40 by return springs 44 and 45 in ten-
sion and pressure force 66 exerted on the face of
piston 42 by the compressed gas in chamber 39, positive
15.
1053733
displacement 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 "hottom
dead point" time flux 62 in magnetic circuit 30 is
less than half of its maximum value.
Magnetic force 64 will continue to decline
after bottom dead point time 265 degrees so ~hat
spring force 65 and pressure force 66 govern move-
ment of the armature and piston and return themoving assembly in the positive displacement or re-
turn direction. Winding current 61 reaches a zero
value at time 320 degrees. Because the current is
rectified, voltage 60 drops 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 pressure force 66 and spring
force 65. Positive displacement 63 will continue
to increase until the moving assembly reaches its
top dead point of operation under the influence
of the spring and magnetic forces as outlined above.
16.
lV53733 ` ~
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 at which its top
5 and bottom end faces are flush with the faces defining
the axially opposite ends of the air-gap. ~t this
point, the armature completely fills the air-gap and,
since the air space between the armature and pole
faces is constant, the reluctance of the total mag-
netic circuit is at a minimum. The armature can thusbe 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, pursuant to the present invention,
it has been discovered that maximum compressor effi-
ciency is obtained when the conical armature "fills
the air-gap" at the point of maximum flux and that,
since this maximum flux point will not necessarily
occur at the "bottom dead point" of operation, it is
advantageous to have the armature continue through the
air-gap beyond this flush point. Since the air space
between the armature and pole faces is no longer con-
stant and is, in fact, a function of axial displacement,the reluctance of the total magnetic circuit will con-
tinue to decrease even though part of the armature is
moving out of the air-gap.
17.
~05~
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 l'b" beyond the end
face of cylinder block 38. The flush condition will
exist when the minimum diameters of the armature and
air-gap are coplanar--i.e., when a = 0. It is un-
desirable 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
pole faces.
Referring to FIG. 6, it will be seen that
the "bottom dead point" of operation is achieved at
an electrical time of about 265 degrees. At this
time flux 62 in magnetic circuit 30 is less than half
of its maximum value. Armature 40 is to be positioned
on rod 41 so that the arma~ure is flush with pole
pieces 32a and 32b at an electrical time of approxi-
mately 180, at which time flux 62 achieves its maxi-
mum 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 shown in FIG. 6. Referring
18.
~1~53733
again to FIG. 4, in the 450 sTu/Hr working embodiment
of the invention armature 40 is positioned to extend
approximately .350 inches beyond pole pieces when
piston 42 extends .030 inches beyond the head-end of
cylinder block 38.
The 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 is optimized when the dis-
tance between the piston face and the valve plate
approaches zero at the "bottom dead point" of opera-
tion. Magnetic force, spring force and compression
force must be thus optimized to achieve maximum com-
pression efficiency without allowing the piston tostrike the valve plate.
While it has been stated for purposes of
explaining the operation of the invention that the
armature moves "into" and "out of" the air-gap, it
should be noted that the present invention, utilizing
the discovery outlined above, need not move the arma-
ture "entirely out of the air-gap" nor locate "a
major portion thereof" outside of the air-gap at the
19 .
1053733
"top dead point" of operation, contrary to the dis-
closure in the United States Barthalon patents 3,542,495
and 3,461,806 respectively. Indeed, in the embodiment
disclosed 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 charac-
teristics detailed above have been defined - i.e., a
substantially linear reluctance curve over the enti.re
stroke length and an armature flush condition at the
time of rnaximum 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 recited conditions, a piston extension dimen-
sion "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 is preferred.
The aforementioned Barthalon patent teaches
that the efficiency of a linear motor will be optimized
if the reluctance of the magnetic circuit varies lin-
early with armature movement. Pursuant to the presentinvention, it has been discovered that the stahility
of a pump which may occasionally operate below atmos-
pheric pressure, such as a refrigeration compressor,
20.
11)53733
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 of taper of the armature identical to
that of the air-gap, it has been found that this con-
dition gives the best overall results. It has 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 armature and
perpendicular to the flux through the pole piece and
should be substantially square rather than rectangular
to achieve the minimum winding length per unit of de-
sired flux. The gross cross sectional area of thearmature, that is, the cross sectional area of the
armature taken as above but including the armature
rod, shoul~ be greater than the effective cross sec-
tional area of the pole piece. This arrangement yields
good results, particularly when an armature rod of
magnetically permeable material is used to increase the
"magnetic cross section" of the armature.
~05~733
The reluctance 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 inches of dis-
placement as measured from the condition of FIG. 4
when the armature is seated in the magnetic core. The
ordinate measurement of reluctance indicates that mini-
mum 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 mechani-
cal efficiency and a reduced range of conditions for
successful operation. It will he noted that reluc-
tance curve 70 is substantially linear over the entire
stroke of .8 inches and has a slope of approximately
.022 ampere-turns per maxwell-inch. The parameters
of this 450 BTU/Hr working emhodiment which contribute
to this low-sloped, linear reluctance curve, and the
consequent high compressor efficiency, are set forth
in the discussion of the working embodiment detailed
hereinafter.
As stated above, the compression volume
bett~een the piston face and valve plate assembly
should approach zero at the "bottom dead point" of
operation. Since the motor does not contain means to
positively stop movement of the piston in the direc-
tion of compression, it may be expected that the pis-
ton will occasionally strike the face of the valve
~V53~33
plate assembly, thus tendiny to jerk armature 40
along the rod 41 in the direction of the compression
chamber. If the armature is allowed to move in response
to this jerking action, it may be expected that the
armature will eventually strike the face of the pole
pieces, thus damaging the core and armature and causing
loud acoustical noise as well as detuning the mechanism.
It is, therefore, desirable to inhihit movement of the
armature with respect to the piston. ,`~eans for accom-
plishing such a purpose is shown in FIG. 8. Piston
80 contains head portion 81 and shank portion 82 ex-
tending along rod 41 into abutment with armature 40.
With piston 80 secured to rod 41, as by press fit,
shrink fit, adhesive and/or being made integral, and
also abutting armature 40, movement of the armature
with respect to the piston in response to the jerking
action above mentioned is prevented.
FIG. 9 shows an alternative to the three-
turn paired spring arrangement in compressor 20 of
FIG. 1. A pair of single-turn linear springs 90 and
91 extend between bracket 46 and clamp 50 through a
slotted outboard hearing and spring retainer plate 92.
Springs 90 and 91 are clamped at tangs 90a, 91a and
90b, 91b to bracket 46 and clamp 50 respectively.
Plate 92 is slotted whexe springs 90 and 91 pass
therethrough so that movement of armature 40 toward
valve plate 51, which movement results in a linear
extension of the single coil of sprinys 90 and 91, will
~053733
not cause interference between the sprinys and plate
92. The use of single-turn springs 90 and 91 reduces
the overall axial dimension of motor 20 and also re-
duces the twisting forces imparted upon armature 40
by the springs during axial reciprocation of the arma-
ture.
A second alternative to the three-turn paired
spring arrangement in the compressor of FIG. 1 is
shown in FIGS. 10 and 11. An outboard bearing and
spring retainer plate 100 is clamped to magnetic cir-
cuit 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 terminations 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 asso-
ciated 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 springs, a "zero pitch" spring
will have a pitch between zero and the diameter of the
spring material. Where straight end-tangs are re-
quired, the spring is usually first coiled on a cir-
cular mandrel or jig with the end-tangs extending
24.
1053733
tangentially from the coil. The end-tangs are then
bent to positions perpendicular from the plane of the
coil. The pitch of the spring thus formed will be sub-
stantially equal to zero within some tolerance range
which depends upon the resilience 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 spring 110 are thus laterally spaced from each
other allowing room for spring 112 to pass therethrough
before terminating in clamp 50. Similarly, end-tangs
112a and 112b are spaced to allow passage of spring 110
therebetween, thereby internesting the springs. In
this geometry the coils of springs 110 and 112 are
aligned with a line connecting tie bolts 102 and 104
rather than being perpendicular therewith and are con-
tained within the lateral perimeter of compressor 20
defined by magnetic circuit 30, thereby reducing the
lateral and axial dimensions of the compressor. Fur-
thermore, with the coils of springs 110 and 112 disposed
in axial proximity to magnetic circuit 30, housing 21
which encompasses compressor 20 may assume a eliptical
shape which is believed to reduce the level of acousti-
cal noise eminating from an operating unit.
The zero pitch internesting springs shown in
FIGS. 10 and 11 have the additional advantage of re-
ducing the twisting forces imparted upon armature 40
almost to zero. This reduction in the torsion or
- 25 -
~L05;~733
twisting forces on the armature and springs results i3 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. Alternatively, 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 s~me magnetic strip material. Such a
core 120 is shown in FIG. 12 and is constructed by 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 strip material
i5 to be tacked as at 130, but need not be terminated. Loops
122 and 124 are then placed in flat end abutment on plane 129
and the magnetic strip material extending from tack 130, or
a separate strip material tacked onto either loop at a conven-
ient attachment point, is wound around the exposed 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, Edwin A. Link, dated 18th
November 1947. 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
~o53t733
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 mag-
netic core shown in FIG. 12 is more easily assembled
and has less waste material than stacked lamination
core 31.
FIG. 13 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 less than the width
of the core (i.e., the dimension 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 connections 140 and 142 on either
side of the machined gap. In order to mount windings
32 and 33 upon pole pieces 136 and 138, connections 140
and 1~2 must be removed at a second machining stage in
the fabrication of magnetic core 120.
FIG~ 14 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 across the faces of pole pieces 136 and 138
in a pair of planes X and Y respectively perpendicular
to the central axis of poles 136 and 138 and parallel
to but displaced on opposite sides of the axis of re-
ciprocation. 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 opposing faces of pole pieces 136 and 138
S3733
respectively. When the yap 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
S sufficient.
However, it has been discovered pursuant to
the present invention that compressor operation is
enhanced when the distance between planes X and Y is
increased beyond this 16.7 per cent figure. In a
specific 450 ~TU/Hr working embodiment of the present
invention having 1.5 inch-square poles, the distance
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
results shown in FIG. 15 when compared to a similar
450 BTU/Hr unit with a planar gap of .375 inches or
16.7 per cent. In FIG. 15 ~TU/Hr output is plotted
versus evaporation temperature. Dashed curve 150
depicts the output of the .375 inch unit over a wide
range of evaporation temperatures while curve 152
represents the output of the .8 inch unit over the
same range. It can be seen tnat the two units per-
form equally at rating point conditions - point 154 -
and perform similarly at evaporation temperatures
lower than rating point. However, at higher evapora-
tion temperatures the per-formance of the .375 inch
unit falls off much more rapidly than the performance
of the .8 inch unit. It should be noted that the
curves of FIG. 15 were plotted from actual test results
and are to scale.
28.
1053733
Strip wound core 120 may replace laminated cor~ 31
in compressor 20 of FIG. 2. In this compressor assembly, tie
bolts 35 and 36 pass tllrough a pair of substantially triangular
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 best seen in FIG. 12. ~pertures 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,
the strip wound core is not compressible in the direction of
tightening of tie bolts 35 and 36. For this reason, it is
easier to hold alignmant tolerances when core 120 is used.
The material disclosed immediately above with refer-
ence to FIGS. 10-12 is the subject of a separate United States
patent application of Richard A. Stuber having the same filing
date as the subject application and assigned to the assignee
of this application. The general concept of tailoring the pole
pieces 136, 138 such as along planes X and Y and shown in FIG. 14
and its effect on performance is part of the present invention.
The specific air-gap sequentially fabricated as shown in FIGS.
13 and 14 and as described above in connection with the specific
ratios of plane spacing to pole cross section to achieve the
result graphically depicted in FIG. 15 is, however, the subject
of the above-mentioned application of Richard ~. Stuber. These
concepts are disclosed herein as being part of the best method
presently known for practicing the subject invention.
Pursuant to the present invention, several working
embodiments of compressor 20 have been built and tested; one
such embodiment is the 450 BTU/Hr (nominal) unit mentioned above
and drawn to scale in FIGS. 1-5. By way of example and not by
way of limitation, the parameters which contribute to the low-
slope linear reluctance curve and the resulting high compressor
29.
~053733
efficiency at rating point conditions are
as follows:
mass of piston 42 . . . . . . . . . . . . . 0.17 lbm
mass of armature 40 . . . . . . . . . . . . 0.8 lbl~
mass of rod 41. . . . . . . . . . . . . . . 0.13 lbm
mass of clamp 50. . . . . . . . . . . . . . 0.12 lbm
effective mass of springs 44
and 45 (1/3 actual mass). . . . . . . . . 0.08 lbm
rate of springs 44 and 45 . . . . . . . . . 200 lb/in
material of rod 41 . . . . . . . . . . . . 1060 steel
net cross-sectional 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.
resistance 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
refrigerant temperature entering
compressor 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 point conditions
after 10,000 hours of operation:
30.
1053733
capacity. . . . . . . 485 BTU/Hr
power input . . . . . 13~ watts
efficiency . . . . . 3.62 s~ru/watt-hour (Weston)
In addition, the following results, which are diffi-
cult to accurately measure in a working linear com-
pressor, were calculated from a computer analysis of
the 450 sT~/Hr model, the a~alysis being similar to
that set forth above with reference to FIG. 6:
length of stroke . . . . . . . . . . Ø54 in
position of A/C power cycle at
"top dead point" of operation. . . . 91 degrees
position of A/C power cycle at
flush position . . . . . . . . . . . 2~7 degrees
current at flush position. . . . . . . 4.9 amps
flux at flush position . . . . . . . . 213 kilomaxwells
position of A/C power at
maximum flux . . . . . . . . . . . . 180 degraess
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
position of A/C power at
opening of discharge valve . . . . . 252 degrees
As discussed above, reluctance curve 70 at FIG. 7
indicates that this embodiment achieved the objective
of having a low-sloped, linear reluctance curve. Fut-
thermore, the above data indicatas that the objective
of achieving maximum flux at the flush position has
been achieved within 8 per cent.
~OS3733
Further embodiments of the present invention
having greater or lesser pumping capacities may be
constructed using the parameters set forth above with
respect to the 450 BTU per hour embodiment of the in-
vention by using the following equations:
1. E sin(wt) - S [10 8N2] di + S k - 10 8N dR dx] i
[ RS ] dt dx dt]
2. M [ d2x] S + C [ dx ] S + KxS = i2 [ (-4.4)10 8N2 dR] S + SF
[ dt2~ [ dt ] [ R2S dx]
3. Wn = ~ 386 (K + K~) (M + m)
here
C = coefficient of viscos friction
E = Zero to peak maximum sine wave voltage
F = pressure force on piston
i = instantaneous current
K = spring rate of return means
Kp = average pneumatic spring constant
m = mass of rod, armature and piston
assembly along with a lumped spring
contribution
~1 = gross compressor assembly wt minus (m)
N = number of turns in the motor windings
r = motor winding resistance
R = magnetic reluctance
S = scaling factor
t = instantaneous value for time
x = displacement of the armature and piston
Wm = natural mechanical frequency of the
mechanical elements associated with m
w = frequency of the electrical circuit
32.
~053733
The above equations may ~e readily derived from well-
known art in the field of mechanical dynamics and elec-
tromagnetics as applied to a linear compressor. The
basis for equation 1 is found in Roters, Herbert C.,
"Electromagnetic Devices," 1st ed., Wiley, `L~ew York,
1963. The basis for equation 2 will be found in Shames,
Irvillg A., "Engineering ~5echanics-~ynamics." 2nd ed.,
Prentice Hall, ~nglewood Cliffs, New Jersey, 1966. The
scaling factor S has been introduced into equations 1
and ~ to facilitate the development of compressors with
equivalent performance over a range of capacities. The
final equation 3 is from the above-mentioned Curwen
article.
It has been discovered, pursuant to the
present invention, that, by mathematically inserting
scaling factor S into the above equations, th~se equa-
tions may be used to approximate the dimensions and
parameters of alternative embodiments of the present
invention. More specifically, the parameters of a
scaled embodiment may be derived from those of an oper-
ative embodiment as follows:
New capacity equals reference X S
~ew bore equals reference x v~
New moving mass equals reference X S
New spring rate equals reference ~ S
New flux path area equals reference X
New coil resistance equals reference X l/S
New coil turns equals reference X l/~r~
where S is a positive real number. ~s stated above,
application of scaling factor S to the parameters of
~053733
the reference embodiment will resul-t in approximate
dimensions and parameters for the alternative embodi-
ment. Translation of these approximate dimensions and
parameters in~o a working model may require some minor
parameter adjustments in the directly scaled replica to
achieve the most efficient combination of parameters,
but such empirical adjustments are believed to be well
within the ordinary skill in the art and do not negate
the substantial savings resulting from application of
these scaling principles.
The disclosed 450 BTU per hour working embodi-
ment of the present invention was scaled from an earlier
embodiment by application of a scaling factor of 1.33
to the parameters of the earlier embodiment. The re-
sulting 450 BTU per hour embodiment had substantially
the same stroke efficiency and performance characteris-
tics as the reference embodiment and performed substan-
tially as predicted.
From the foregoing description, it will now be
apparent that there has been provided, in accordance with
the invention, a moving-iron linear compressor motor
that fully satisfies the objects and advantages set forth
above. While the invention has been described in con-
junction with specific embodiments thereof, it is evi-
dent that many alternatives, modifications, and varia-
tions 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
- 34 -
10~;~733
refrigerant systems and that, indeed, many principles of the
invention may be applied generally to gas pumps, such as air
compressors or the like.
This application is a divisional application from
copending Canadian application serial no. 235310 filed on 12th
September 1975.
- 35--
