TECHNICAL FIELD
[0001] The present invention relates to rotary compressors, and particularly relates to
measures against vibration caused by a variation in load torque.
BACKGROUND ART
[0002] As a conventional rotary compressor including two cylinder chambers, a compressor
which compresses refrigerant by utilizing a variation in the volume of a cylinder
chamber caused by eccentric rotation of an annular piston has been employed, as disclosed
in, for example, Patent Document 1.
[0003] The compressor of Patent Document 1 includes: a cylinder having an annular cylinder
chamber; and an annular piston placed in the cylinder chamber. The cylinder is composed
of concentrically disposed outer and inner cylinders. Specifically, a cylinder chamber
is formed between the outer cylinder and the inner cylinder, and this cylinder chamber
is partitioned into an outer cylinder chamber and an inner cylinder chamber with an
annular piston. The annular piston is configured to eccentrically rotate about the
cylinder center by driving an electric motor, with the outer peripheral surface of
the annular piston being in contact with the inner peripheral surface of the outer
cylinder at substantially one point, and with the inner peripheral surface of the
annular piston being in contact with the outer peripheral surface of the inner cylinder
at substantially one point.
[0004] An outer blade is provided outside the annular piston. An inner blade is provided
inside the annular piston on an extension of the outer blade. The outer blade is inserted
in the outer cylinder, is radially biased toward the inside of the annular piston,
and has its tip in pressure contact with the outer peripheral surface of the annular
piston. The inner blade is inserted in the inner cylinder, is radially biased toward
the outside of the annular piston, and has its tip in pressure contact with the inner
peripheral surface of the annular piston. The outer and inner blades respectively
partition the outer and inner cylinder chambers into high-pressure chambers and low-pressure
chambers. In this compressor, eccentric rotation of the annular piston causes fluid
to be sucked into the low-pressure chamber and to be compressed in the high-pressure
chamber in each of the cylinder chambers.
Patent Document 1: Japanese Laid-Open Patent Publication No. 6-288358
DISCLOSURE OF INVENTION
Problems that the Invention is to Solve
[0005] However, disadvantageously, in the compressor of the Patent Document 1, the load
torque of a drive shaft varies in one turn of rotation. This variation causes the
rotation speeds of the drive shaft and a rotor of the electric motor for driving the
drive shaft to vary, resulting in that vibration occurs in the tangential direction
of a casing in which a stator of the electric motor is fixed. In a possible worst
case, a pipe connected to the casing might be broken.
[0006] It is therefore an object of the present invention to suppress vibration caused by
a load torque variation in one turn of rotation in a rotary compressor in which relative
eccentric rotation is accomplished between a piston and a cylinder having two cylinder
chambers.
Means of Solving the Problems
[0007] A first aspect of the present invention is directed to a rotary compressor including:
a compression mechanism (20, 80) including a piston (22, 87a, 87b) and a cylinder
(21, 81a, 81b) having two cylinder chambers (C1, C2, 82a, 82b); and an electric motor
(30, 65) configured to change volumes of the cylinder chambers (C1, C2, 82a, 82b)
by causing relative eccentric rotation between the cylinder (21, 81a, 81b) and the
piston (22, 87a, 87b). The rotary compressor further includes a torque control means
(50) configured to change an output torque of the electric motor (30, 65) in accordance
with a variation in a load torque of the compression mechanism (20, 80) in one turn
of rotation.
[0008] In this aspect, the relative eccentric rotation of the cylinder (21, 81a, 81b) and
the piston (22, 87a, 87b) caused by driving the electric motor (30, 65) causes the
volumes of the two cylinder chambers (C1, C2, 82a, 82b) to vary. In each of the cylinder
chambers (C1, C2, 82a, 82b), fluid is sucked as the volume of the low-pressure chamber
(i.e., the suction chamber) increases, whereas fluid in the high-pressure chamber
(i.e., the compression chamber) is compressed as the volume of the high-pressure chamber
decreases.
[0009] The load torque of the electric motor (30, 65) varies in accordance with the rotation
angle in one turn of rotation of the compression mechanism (20, 80). Specifically,
in each of the cylinder chambers (C1, C2, 82a, 82b), the load torque is highest substantially
immediately before and after compressed fluid starts to be discharged. Accordingly,
in this state, the rotation speed of the cylinder (21, 81a, 81b) or the piston (22,
87a, 87b) varies because the output torque of the electric motor (30, 65) is fixed.
That is, when the load torque increases, the rotation speed decreases, whereas when
the load torque decreases, the rotation speed increases. This variation in the rotation
speed causes vibration of the compressor in the tangential direction of the casing.
[0010] On the other hand, in this aspect of the present invention, the torque control means
(50) adjusts the output torque of the electric motor (30, 65) in accordance with a
variation in the load torque in one turn of rotation. Specifically, the output torque
of the electric motor (30, 65) decreases as the load torque decreases, and increases
as the load torque increases. That is, torque control is performed such that the output
torque of the electric motor (30, 65) is adjusted to a value commensurate with the
load torque. This control makes the rotation speed of the cylinder (21, 81a, 81b)
or the piston (22, 87a, 87b) substantially constant, thus suppressing vibration of
the compressor.
[0011] In a second aspect of the present invention, in the rotary compressor of the first
aspect, the cylinder (21) includes an annular cylinder chamber (C1, C2), and the piston
(22) is an annular piston (22) housed in the annular cylinder chamber (C1, C2) and
partitioning the annular cylinder chamber (C1, C2) into two cylinder chambers which
are an outer cylinder chamber (C1) and an inner cylinder chamber (C2).
[0012] In this aspect, relative eccentric rotation of the cylinder (21) and the annular
piston (22 (52)) caused by driving the electric motor (30) causes the volumes of the
outer cylinder chamber (C1) and the inner cylinder chamber (C2) to vary. In each of
the cylinder chambers (C1, C2), fluid is sucked as the volume of the low-pressure
chamber (i.e., the suction chamber) increases, whereas fluid in the high-pressure
chamber (i.e., the compression chamber) is compressed as the volume of the high-pressure
chamber decreases.
[0013] In this case, the load torque of the electric motor (30) also varies in accordance
with the rotation angle in one turn of rotation of the compression mechanism (20),
thereby causing the rotation speed of the cylinder (21) or the annular piston (22)
to vary. This variation causes vibration of the compressor in the tangential direction
of the casing. On the other hand, in the present invention, the torque control means
(50) adjusts the output torque of the electric motor (30) in accordance with a variation
in the load torque in one turn of rotation. Accordingly, the rotation speed of the
cylinder (21) or the annular piston (22) becomes substantially constant, thus suppressing
vibration of the compressor.
[0014] In a third aspect of the present invention, in the rotary compressor of the second
aspect, a volume ratio of the inner cylinder chamber (C2) to the outer cylinder chamber
(C1) is in the range from 0.6 to 1.0.
[0015] In this aspect, since the volume ratio of the inner cylinder chamber (C2) to the
outer cylinder chamber (C1) is set in the range from 0.6 to 1.0, the variation range
in the load torque in one turn is small. Specifically, as shown in FIG. 5, as the
volume ratio Vr of the inner cylinder chamber (C2) to the outer cylinder chamber (C1)
decreases, the variation range (i.e., the amount of variation) in the load torque
increases. In particular, when the volume ratio Vr is about 0.6 or less, the variation
range in the load torque becomes extremely large.
[0016] The torque control of the electric motor (30) changes the output torque of the electric
motor (30) by adjusting, for example, an input current or an input voltage to the
electric motor (30). For example, when the load torque is high, the input current
is increased so as to increase the output torque of the electric motor (30). On the
other hand, when the load torque is low, the input current is reduced so as to reduce
the output torque of the electric motor (30). In general, the electric motor (30)
exhibits a high operational efficiency when the electric motor (30) is driven with
a substantially constant input current or voltage. Specifically, when the amount of
variation (i.e., the degree of control) in, for example, the input current becomes
large, the operational efficiency of the electric motor (30) greatly decreases.
[0017] On the other hand, in the present invention, the volume ratio Vr of the inner cylinder
chamber (C2) to the outer cylinder chamber (C1) is limited within a given range as
described above, thus reducing the amount of variation in the load torque in one turn
of rotation. Accordingly, the amount of variation in the input current or the input
voltage of the electric motor (30) is reduced in one turn of rotation, resulting in
suppressing a decrease in the operational efficiency of the electric motor (30).
[0018] In a fourth aspect of the present invention, in the rotary compressor of the second
or third aspect, the electric motor (30) is a brushless DC motor.
[0019] In this aspect, since a brushless direct-current (DC) motor is used as the electric
motor (30), the operational efficiency of the electric motor (30) is higher than in
the case of using an alternating-current (AC) motor. In particular, in torque control
performed during low-speed rotation in which variation in the rotation speed is likely
to be large, the DC motor can maintain a relatively high efficiently to a low speed,
although the AC motor greatly decreases in efficiency, and thus substantially is not
operable.
[0020] In a fifth aspect of the present invention, in the rotary compressor of the second
or third aspect, the torque control means (50) is configured to change an output torque
of the electric motor (30) by changing one of an input current, an input voltage,
and an input current phase of the electric motor (30).
[0021] In this aspect, when the load torque decreases in one turn of rotation, the input
current or the input voltage is reduced, thereby reducing the output torque of the
electric motor (30). When the load torque increases, the input current or the input
voltage is increased, thereby increasing the output torque of the electric motor (30).
In this manner, the output torque of the electric motor (30) is adjusted to a value
commensurate with the load torque. In addition, by adjusting (i.e., moving forward
or backward) the input current phase, the output torque of the electric motor (30)
increases or decreases to a value commensurate with the load torque. In particular,
this adjustment of the input current phase can allow the output torque of the electric
motor (30) to more closely follow a load torque which abruptly changes.
[0022] In a sixth aspect of the present invention, in the rotary compressor of the second
or third aspect, the electric motor (30) is coupled to the cylinder (21) configured
to rotate relative to the annular piston (22) which is stationary.
[0023] In this aspect, the cylinder (21) is movable, and the annular piston (22) is stationary.
Specifically, the cylinder (21) eccentrically rotates relative to the annular piston
(22), and the torque control means (50) suppresses a variation in the rotation speed
of the cylinder (21). As a result, vibration caused by the variation in the rotation
speed of the cylinder (21) can be suppressed.
[0024] In a seventh aspect of the present invention, the rotary compressor of the second
or third aspect, the electric motor (30) is coupled to the annular piston (22) configured
to rotate relative to the cylinder (21) which is stationary.
[0025] In this aspect, the cylinder (21) is stationary, and the annular piston (52) is movable.
Specifically, the annular piston (22) eccentrically rotates relative to the cylinder
(21), and the torque control means (50) suppresses a variation in the rotation speed
of the annular piston (22). As a result, vibration caused by the variation in the
rotation speed of the annular piston (22) can be suppressed.
[0026] In an eighth aspect of the present invention, the rotary compressor of the second
aspect, the compression mechanism (20) is configured to perform two-stage compression
on fluid, with one of the outer and inner cylinder chambers (C1) and (C2) used at
a low-stage side and the other one of the outer and inner cylinder chambers (C1) and
(C2) used at a high-stage side.
[0027] In this aspect, first, low-pressure fluid sucked into the outer cylinder chamber
(C1) is compressed to be intermediate-pressure fluid. This intermediate-pressure fluid
is sucked into the inner cylinder chamber (C2). This intermediate-pressure fluid in
the inner cylinder chamber (C2) is further compressed to be high-pressure fluid. This
series of processes is performed in one turn of rotation of the compression mechanism
(20), thus changing the load torque of the electric motor (30) in accordance with
the rotation angle. In this case, the torque control means (50) also makes the rotation
speed of the cylinder (21) or the annular piston (22) substantially constant, resulting
in suppressing vibration of the compressor.
[0028] In a ninth aspect of the present invention, the rotary compressor of the eighth aspect,
a volume ratio of the inner cylinder chamber (C2) to the outer cylinder chamber (C1)
is in the range from 0.6 to 0.8.
[0029] In this aspect, the volume ratio of the inner cylinder chamber (C2) to the outer
cylinder chamber (C1) is set in the range from 0.6 to 0.8, and thus the variation
range in the load torque in one turn of rotation is small. Specifically, as shown
in FIG. 13, the variation range (i.e., the amount of variation) in the load torque
is smaller in cases where the volume ratio Vr of the inner cylinder chamber (C2) to
the outer cylinder chamber (C1) is 0.6 and 0.8, than in a case where the volume ratio
Vr is 0.5 (or 1.0). In the case of the volume ratio Vr=1.0, the outer cylinder chamber
(C1) and the inner cylinder chamber (C2) have the same volume, and thus this case
corresponds to a one-cylinder compression mechanism performing so-called single-stage
compression. In this aspect, the variation range in the load torque in one turn of
rotation is smaller than in the one-cylinder compression mechanism.
[0030] In a tenth aspect of the present invention, in the rotary compressor of the first
aspect, the cylinders (81 a, 81b) are respectively a low-stage first cylinder (81
a) and a high-stage second cylinder (81b) both including cylinder chambers (82a, 82b),
the pistons (87a, 87b) are respectively a first rotary piston (87a) housed in the
cylinder chamber (82a) of the first cylinder (81a) and a second rotary piston (87b)
housed in the cylinder chamber (82b) of the second cylinder (81b), and the compression
mechanism (80) is configured to perform two-stage compression on fluid in the cylinders
(81a, 81b) by eccentric rotation of the rotary pistons (87a, 87b) caused by the electric
motor (65).
[0031] In this aspect, the compression mechanism (80) is a so-called two-cylinder rotary
compression mechanism. In this compression mechanism (80), the rotary pistons (87a,
87b) eccentrically rotate by driving the electric motor (65). This eccentric rotation
of the rotary pistons (87a, 87b) causes the volumes of the cylinder chambers (82a,
82b) to vary, thereby achieving two-compression of fluid in the cylinders (81a, 81b).
Specifically, first, low-pressure fluid sucked into the cylinder chamber (82a) of
the first cylinder (81a) is compressed to be intermediate-pressure fluid. This intermediate-pressure
fluid is sucked into the cylinder chamber (82b) of the second cylinder chamber (82b).
The intermediate-pressure fluid in the cylinder chamber (82b) is further compressed
to be high-pressure fluid. This series of processes is performed in one turn of rotation
of the compression mechanism (20), thus changing the load torque of the electric motor
(30) in accordance with the rotation angle. In this case, the torque control means
(50) also makes the rotation speed of the rotary pistons (87a, 87b) substantially
constant, resulting in suppressing vibration of the compressor.
[0032] In an eleventh aspect of the present invention, in the rotary compressor of the tenth
aspect, a volume ratio of the cylinder chamber (82b) of the second cylinder (81b)
to the cylinder chamber (82a) of the first cylinder (8 1 a) is in the range from 0.6
to 0.8.
[0033] In this aspect, the volume ratio of the cylinder chamber (82b) of the second cylinder
(81b) to the cylinder chamber (82a) of the first cylinder (81a) is set in the range
from 0.6 to 0.8, and thus the variation range in the load torque in one turn of rotation
is small. Specifically, as shown in FIG. 13, the variation range (i.e., the amount
of variation) in the load torque is smaller in cases where the volume ratio Vr of
the inner cylinder chamber (C2) to the outer cylinder chamber (C1) is 0.6 and 0.8,
than in a case where the volume ratio Vr is 0.5 (or 1.0).
[0034] In a twelfth aspect of the present invention, in the rotary compressor of the tenth
or eleventh aspect, the compression mechanism (80) is configured such that a rotational
phase of the rotary piston (87a) of the first cylinder (81a) is shifted from a rotational
phase of the rotary piston (87b) of the second cylinder (81 b) by 180°.
[0035] In this aspect, when the volume of the cylinder chamber (82a) decreases in the first
cylinder (81a) with the rotation of the rotary piston (87a), fluid compressed to an
intermediate pressure is discharged. At substantially the same time, the volume of
the cylinder chamber (82b) increases in the second cylinder (81b) with the rotation
of the rotary piston (87b), thereby sucking the intermediate-pressure fluid discharged
from the first cylinder (81a). This sucked intermediate-pressure fluid is further
compressed with a decrease in the volume of the cylinder chamber (82b) of the second
cylinder (81b), and then is discharged.
EFFECTS OF THE INVENTION
[0036] Accordingly, in the present invention, the output torque of the electric motor (30,
65) is adjusted according to a variation in the load torque of the compression mechanism
(20, 80) in one turn of rotation. Thus, a variation in the rotation speed of the cylinder
(21, 81a, 81b) or the piston (22, 87a, 87b) can be reduced. As a result, vibration
of the compressor caused by a variation in the rotation speed can be suppressed.
[0037] In the third aspect of the present invention, the volume ratio of the inner cylinder
chamber (C2) to the outer cylinder chamber (C1) is limited within a given range (i.e.,
from 0.6 to 1.0), and thus the amount of variation in the load torque in one turn
of rotation can be reduced. Accordingly, the amount of variation in the output torque
of the electric motor (30) can be reduced, thus suppressing a decrease in the operational
efficiency of the electric motor (30). As a result, energy saving in operating the
compressor can be achieved.
[0038] In the fourth aspect of the present invention, a brushless direct-current motor is
used as the electric motor (30). Thus, the efficiency of the electric motor (30) can
be enhanced, as compared to the case of using an AC motor. As a result, further energy
saving in operating the compressor can be achieved.
[0039] In the ninth or eleventh aspect of the present invention, the volume ratio of the
high-stage cylinder chamber (C2, 82b) to the low-stage cylinder chamber (C1, 82a)
is also limited within a given range (i.e., from 0.6 to 0.8) in a two-cylinder two-stage
compression mechanism. Accordingly, the amount of variation in the load torque in
one turn of rotation can be reduced, thus suppressing a decrease in the operational
efficiency of the electric motor (30, 65).
BRIEF DESCRIPTION OF DRAWINGS
[0040]
[FIG. 1] FIG. 1 is a longitudinal cross-sectional view illustrating a compressor according
to a first embodiment.
[FIG. 2] FIG. 2 is a transverse cross-sectional view illustrating a compression mechanism
of the first embodiment.
[FIG. 3] FIG. 3 shows transverse cross sections showing operation of the compression
mechanism of the first embodiment for every 90° of rotation.
[FIG. 4] FIG. 4 is a graph showing variations in compression torques in one turn of
rotation.
[FIG. 5] FIG. 5 is a graph showing variations in compression torques with respect
to the volume ratios Vr.
[FIG. 6] FIG. 6 is a graph showing the amounts of decrease in a torque variation ratio,
a vibration ratio, and a motor efficiency with respect to volume ratios Vr.
[FIG. 7] FIG. 7 is a longitudinal cross-sectional view illustrating a compressor according
to a second embodiment.
[FIG. 8] FIG. 8 is a transverse cross-sectional view illustrating a compression mechanism
of the second embodiment.
[FIG. 9] FIG. 9 shows transverse cross-sections showing operation of the compression
mechanism of the second embodiment for every 90° of rotation.
[FIG. 10] FIG. 10 is a longitudinal cross-sectional view illustrating a compressor
according to a third embodiment.
[FIG. 11] FIG. 11 is a transverse cross-sectional view illustrating a compression
mechanism of the third embodiment.
[FIG. 12] FIG. 12 is a graph showing relationships between operating pressure ratios
and compression efficiencies based on volume ratios Vr.
[FIG. 13] FIG. 13 is a graph showing variations in compression torque with respect
to volume ratios Vr.
[FIG. 14] FIG. 14 is a graph showing a torque variation ratio with respect to the
volume ratio Vr.
[FIG. 15] FIG. 15 is a longitudinal cross-sectional view illustrating a compressor
according to a fourth embodiment.
[FIG. 16] FIG. 16 is a transverse cross-sectional view illustrating a compression
mechanism of the fourth embodiment.
Description of Characters
[0041]
- 1, 60
- compressor
- 20
- compression mechanism
- 21
- cylinder
- 22
- annular piston (piston)
- 30
- electric motor
- 50
- controller (torque control means)
- C1
- outer cylinder chamber (cylinder chamber)
- C2
- inner cylinder chamber (cylinder chamber)
- 65
- electric motor
- 80
- compression mechanism
- 81a
- first cylinder (cylinder)
- 81b
- second cylinder (cylinder)
- 82a
- first cylinder chamber (cylinder chamber)
- 82b
- second cylinder chamber (cylinder chamber)
- 87a
- first rotary piston (piston)
- 87b
- second rotary piston (piston)
BEST MODE FOR CARRYING OUT THE INVENTION
[0042] Hereinafter, embodiments of the present invention will be specifically described
with reference to the drawings. The following embodiments are merely preferred examples
in nature, and are not intended to limit the scope, applications, and use of the invention.
«EMBODIMENT 1»
[0043] A first embodiment is directed to a rotary compressor as illustrated in FIG. 1. This
compressor (1) is of a fully-enclosed type in which a compression mechanism (20) and
an electric motor (30) for driving the compression mechanism (20) are housed in a
casing (10), and are fully enclosed. The compressor (1) is used for compressing refrigerant
sucked from an evaporator and for discharging the compressed refrigerant into a condenser
in a refrigerant circuit of an air conditioner, for example.
[0044] The casing (10) includes a cylindrical body (11) and upper and lower heads (12) and
(13) respectively fixed to the top and bottom of the body (11). A suction pipe (14)
penetrates the upper head (12), and a discharge pipe (15) penetrates the body (11).
[0045] The compression mechanism (20) includes: upper and lower housings (16) and (17) fixed
to the casing (10); and a cylinder (21). The cylinder (21) has an annular cylinder
chamber (C1, C2), and is located between the upper housing (16) and the lower housing
(17). The upper housing (16) includes an annular piston (22) which is located in the
cylinder chamber (C1, C2) and is continuous to the upper housing (16). The cylinder
(21) is configured to eccentrically rotate relative to the annular piston (22). Specifically,
in this embodiment, the cylinder (21) and the annular piston (22) provide relative
eccentric rotation in which the cylinder (21) is movable and the annular piston (22)
is stationary.
[0046] The electric motor (30) is a brushless direct-current (DC) motor including a stator
(31) and a rotor (32). The stator (31) is located below the compression mechanism
(20), and is fixed to the body (11) of the casing (10). The rotor (32) is coupled
to a drive shaft (33) which rotates together with the rotor (32). The drive shaft
(33) longitudinally penetrates the compression mechanism (20), and has an eccentric
part (33a) located in the cylinder chamber (C1, C2). This eccentric part (33a) has
a diameter greater than the other portion, and is eccentric from the axis of the drive
shaft (33) to a given extent.
[0047] An axially extending oil supply passageway (not shown) is provided in the drive shaft
(33). An oil supply pump (34) is provided at the bottom of the drive shaft (33). This
oil supply pump (34) pumps lubricating oil accumulated in the bottom of the casing
(10), and supplies the oil to a sliding part of the compression mechanism (20) through
the oil supply passageway of the drive shaft (33).
[0048] The cylinder (21) has an outer cylinder portion (24) and an inner cylinder portion
(25). The outer cylinder portion (24) and the inner cylinder portion (25) are annular,
and have the same axis. These outer and inner cylinder portions (24) and (25) are
coupled to each other at an end by a head (26) to be continuous. The annular cylinder
chamber (C1, C2) is formed between the inner peripheral surface of the outer cylinder
portion (24) and the outer peripheral surface of the inner cylinder portion (25).
The eccentric part (33a) of the drive shaft (33) is slidably fit into the inner cylinder
portion (25). The cylinder (21) is made of cast steel or an aluminum alloy, for example.
[0049] Each of the upper housing (16) and the lower housing (17) has a bearing portion (16a,
17a) for rotatably supporting the drive shaft (33). In this manner, the compressor
(1) of this embodiment has a penetration structure in which the drive shaft (33) longitudinally
penetrates the cylinder chamber (C1, C2) and both axial ends of the eccentric part
(33a) are held in the casing (10) with the bearing portions (16a, 17a) interposed
therebetween.
[0050] The outer peripheral surface of the annular piston (22) has a diameter smaller than
that of the inner peripheral surface of the outer cylinder portion (24), and the inner
peripheral surface of the annular piston (22) has a diameter larger than that of the
outer peripheral surface of the inner cylinder portion (25). The annular piston (22)
is eccentrically placed in the annular cylinder chamber (C1, C2) to partition the
cylinder chamber (C1, C2) into an outer cylinder chamber (C1) and an inner cylinder
chamber (C2). Specifically, the outer cylinder chamber (C1) is formed between the
inner peripheral surface of the outer cylinder portion (24) and the outer peripheral
surface of the annular piston (22). The inner cylinder chamber (C2) is formed between
the inner peripheral surface of the annular piston (22) and the outer peripheral surface
of the inner cylinder portion (25). The head (26) of the cylinder (21) serves as a
first block member for blocking one end of the cylinder chamber (C1, C2), and the
upper housing (16) serves as a second block member for blocking the other end of the
cylinder chamber (C1, C2).
[0051] The outer peripheral surface of the annular piston (22) is in contact with the inner
peripheral surface of the outer cylinder portion (24) at substantially one point.
At a position whose phase is 180° shifted from this contact point, the inner peripheral
surface of the annular piston (22) is in contact with the outer peripheral surface
of the inner cylinder portion (25) at substantially one point.
[0052] As illustrated in FIG. 2, the compression mechanism (20) has a blade (23) which partitions
each of the outer cylinder chamber (C1) and the inner cylinder chamber (C2) into a
high-pressure chamber (C1-Hp, C2-Hp) serving as a compression chamber and a low-pressure
chamber (C1-Lp, C2-Lp) serving as a suction chamber. This blade (23) is in the shape
of a rectangular plate which penetrates the annular piston (22) and extends in the
direction along the diameter of the cylinder chamber (C1, C2). Both ends of the blade
(23) are respectively fixed to the inner peripheral surface of the outer cylinder
portion (24) and the outer peripheral surface of the inner cylinder portion (25).
[0053] The annular piston (22) has a C-shape obtained by partially cutting off the annular
piston (22) so as to allow the blade (23) to penetrate therethrough. The cut-off portion
of the annular piston (22) is provided with swing bushings (27). The swing bushings
(27) are constituted by a discharge-side bushing (27A) and a suction-side bushing
(27B). The discharge-side bushing (27A) and the suction-side bushing (27B) are located
toward the high-pressure chamber (C1-Hp, C2-Hp) and the low-pressure chamber (C1-Lp,
C2-Lp), respectively, with the blade (23) sandwiched therebetween.
[0054] The discharge-side bushing (27A) and the suction-side bushing (27B) are approximately
semicircular in cross section, and have their plane surfaces face each other. That
is, a blade groove (28) in which the blade (23) slides is formed between opposing
surfaces of the bushings (27A, 27B). The swing bushings (27) are configured such that
the blade (23) moves forward and backward in the blade groove (28), with the blade
(23) and the cylinder (21) swinging in an integrated manner relative to the annular
piston (22). The bushings (27A, 27B) are not necessarily separated from each other,
and may be partially coupled to each other to be continuous.
[0055] In the compression mechanism (20), with the rotation of the drive shaft (33), the
points of contact on the annular piston (22) with the outer and inner cylinder portions
(24) and (25) sequentially move from the state shown in FIG. 3(A) to the state shown
in FIG. 3(D). Specifically, the rotation of the drive shaft (33) causes the compression
mechanism (20) to revolve about the drive shaft (33), without causing the outer cylinder
portion (24) and the inner cylinder portion (25) to rotate.
[0056] The upper housing (16) has an inlet (41) in the shape of a slot below the suction
pipe (14). This inlet (41) penetrates the upper housing (16) along the axis of the
upper housing (16), and allows the low-pressure chamber (C1-Lp, C2-Lp) of the cylinder
chamber (C1, C2) to communicate with the space (i.e., low-pressure space (S1)) above
the upper housing (16). The outer cylinder portion (24) has a through hole (43) which
allows suction space (42) to communicate with the low-pressure chamber (C1-Lp) of
the outer cylinder chamber (C1). The annular piston (22) has a through hole (44) which
allows the low-pressure chamber (C1-Lp) of the outer cylinder chamber (C1) to communicate
with the low-pressure chamber (C2-Lp) of the inner cylinder chamber (C2).
[0057] The annular piston (22) and the outer cylinder portion (24) preferably have wedge
shapes by chamfering top portions thereof corresponding to the inlet (41) as indicated
by broken lines in FIG. 1. In this case, refrigerant can be efficiently sucked into
the low-pressure chambers (C1-Lp, C2-Lp).
[0058] The housing (16) has two outlets (45). These outlets (45) penetrate the upper housing
(16) along the axis of the upper housing (16). The bottoms of the outlets (45) are
respectively open to the high-pressure chambers (C1-Hp, C2-Hp) of the cylinder chambers
(C1, C2). On the other hand, the tops of the outlets (45) communicate with discharge
space (47) through discharge valves (46) for opening and closing the outlets (45).
[0059] This discharge space (47) is formed between the upper housing (16) and a cover plate
(18). The upper housing (16) and the lower housing (17) are provided with a discharge
passageway (47a) which allows the discharge space (47) to communicate with space (i.e.,
high-pressure space (S2)) below the lower housing (17).
[0060] The lower housing (17) is provided with a seal ring (29). This seal ring (29) is
placed in an annular groove (17b) of the lower housing (17), and is in pressure contact
with the bottom of the head (26) of the cylinder (21). High-pressure lubricating oil
is introduced between the cylinder (21) and the lower housing (17) at the inner side
of the seal ring (29) in the radial direction of the seal ring (29). In this manner,
the seal ring (29) constitutes a compliance mechanism in which an axial clearance
between the bottom surface of the annular piston (22) and the head (26) of the cylinder
(21) is reduced by utilizing the pressure of the lubricating oil.
[0061] In this embodiment, the volume Vout of the outer cylinder chamber (C1) is larger
than the volume Vin of the inner cylinder chamber (C2). Specifically, the volume ratio
Vr (Vin/Vout) of the inner cylinder chamber (C2) to the outer cylinder chamber (C1)
is set at about 0.7. This volume ratio Vr is preferably set in the range from 0.6
to 1.0.
[0062] The compressor (21) includes a controller (50) which is a torque control means for
controlling the output torque of the electric motor (30).
[0063] The controller (50) is configured to change the output torque of the electric motor
(30) in accordance with a variation in the load torque of the compression mechanism
(20) in one turn of rotation. This controller (50) receives the rotation angle of
the rotor (32), and supplies, to the electric motor (30), current having a value which
has been previously set in accordance with the rotation angle of the rotor (32). That
is, the controller (50) changes the output torque of the electric motor (30) by controlling
input current of the electric motor (30). The rotation angle of the rotor (32) is
equal to the rotation angle of the drive shaft (33). As the rotation angle of the
rotor (32), a value detected by a rotation sensor or a value calculated from the induced
voltage or current of the electric motor (30) is used.
[0064] Specifically, input current is increased at a rotation angle with a high load torque,
whereas the input current is reduced at a rotation angle with a low load torque. Since
the output torque of the electric motor (30) is proportional to the input current,
increase/decrease in the input current causes the output torque of the electric motor
(30) to increase/decrease accordingly. In this manner, the output torque of the electric
motor (30) becomes commensurate with the load torque. Thus, a variation in the rotation
speed of the drive shaft (33) in one turn can be suppressed, resulting in suppression
of vibration.
[0065] In the present invention, instead of input current, an input voltage or an input
current phase may be controlled in accordance with the rotation angle of the rotor
(32) so as to change the output torque of the electric motor (30). For example, the
input voltage is increased at a rotation angle with a high load torque, whereas the
input voltage is reduced at a rotation angle with a low load torque. Accordingly,
the output torque of the electric motor (30) increases/decreases in proportion to
the input voltage, and is changed to a value commensurate with the load torque. A
shift of the input voltage phase causes the output torque of the electric motor (30)
to increase/decrease to a value commensurate with the load torque. In particular,
adjusting the input current phase allows the output torque of the electric motor (30)
to more closely follow the load torque which abruptly changes.
-Operation-
[0066] Now, it is described how the compressor (1) operates with reference to the drawings.
[0067] First, when the electric motor (30) is started, rotation of the rotor (32) is conveyed
to the outer cylinder portion (24) and the inner cylinder portion (25) of the compression
mechanism (20) via the drive shaft (33). Accordingly, the blade (23) reciprocates
(i.e., moves forward and backward) between the swing bushings (27), and the blade
(23) and the swing bushings (27) swing in an integrated manner with respect to the
annular piston (22). The outer cylinder portion (24) and the inner cylinder portion
(25) revolve about the annular piston (22) while swinging, thereby causing the compression
mechanism (20) to perform given compression operation.
[0068] Specifically, in the outer cylinder chamber (C1), the volume of the low-pressure
chamber (C1-Lp) is approximately smallest in the state shown in FIG. 3(D). While the
drive shaft (33) rotates in the clockwise direction in the drawings to shift from
the state of FIG. 3(D) to the states of FIGS. 3(A), 3(B), and 3(C) in this order,
the volume of the low-pressure chamber (C1-Lp) increases. With this increase in the
volume of the low-pressure chamber (C1-Lp), refrigerant passes through the suction
pipe (14), the low-pressure space (S1), and the inlet (41) to be sucked into the low-pressure
chamber (C1-Lp). At this time, the refrigerant is not only sucked directly into the
low-pressure chamber (C1-Lp) from the inlet (41), but also partially enters the suction
space (42) from the inlet (41), and then is sucked into the low-pressure chamber (C1-Lp)
through the through hole (43).
[0069] When the drive shaft (33) returns to the state of FIG. 3(D) after one turn of rotation,
the suction of refrigerant into the low-pressure chamber (C1-Lp) is completed. This
low-pressure chamber (C1-Lp) is then changed to a high-pressure chamber (C1-Hp) in
which the refrigerant is compressed, and a new low-pressure chamber (C1-Lp) is created
with the blade (23) sandwiched between the high-pressure chamber (C1-Hp) and the low-pressure
chamber (C1-Lp). Then, while the drive shaft (33) further rotates, suction of refrigerant
is repeated in the low-pressure chamber (C1-Lp), whereas the volume of the high-pressure
chamber (C1-Hp) decreases, thereby compressing the refrigerant in the high-pressure
chamber (C1-Hp). The high-pressure refrigerant in the high-pressure chamber (C1-Hp)
flows from the outlets (45) into the discharge space (47), and then flows into the
high-pressure space (S2) through the discharge passageway (47a).
[0070] In the inner cylinder chamber (C2), the volume of the low-pressure chamber (C2-Lp)
is approximately smallest in the state shown in FIG. 3(B). While the drive shaft (33)
rotates in the clockwise direction in the drawings from the state of FIG. 3(B) to
the states of FIGS. 3(C), 3(D), and 3(A) in this order, the volume of the low-pressure
chamber (C2-Lp) increases. With this increase in the volume of the low-pressure chamber
(C2-Lp), refrigerant passes through the suction pipe (14), the low-pressure space
(S1), and the inlet (41) to be sucked into the low-pressure chamber (C2-Lp). At this
time, the refrigerant is not only sucked directly into the low-pressure chamber (C2-Lp)
from the inlet (41), but also partially enters the suction space (42) from the inlet
(41), and then passes through the through hole (43), the low-pressure chamber (C1-Lp)
of the outer cylinder chamber, and the through hole (44) to be sucked into the low-pressure
chamber (C2-Lp) of the inner cylinder chamber (C2).
[0071] When the drive shaft (33) returns to the state of FIG. 3(B) after one turn of rotation,
the suction of refrigerant into the low-pressure chamber (C2-Lp) is completed. This
low-pressure chamber (C2-Lp) is then changed to a high-pressure chamber (C2-Hp) in
which the refrigerant is compressed, and a new low-pressure chamber (C2-Lp) is created
with the blade (23) sandwiched between the high-pressure chamber (C2-Hp) and the low-pressure
chamber (C2-Lp). Then, while the drive shaft (33) further rotates, suction of refrigerant
is repeated in the low-pressure chamber (C2-Lp), whereas the volume of the high-pressure
chamber (C2-Hp) decreases, thereby compressing refrigerant in the high-pressure chamber
(C2-Hp). The high-pressure refrigerant in the high-pressure chamber (C2-Hp) flows
from the outlets (45) into the discharge space (47), and then flows into the high-pressure
space (S2) through the discharge passageway (47a).
[0072] In this manner, high-pressure refrigerant which has been compressed in the outer
cylinder chamber (C1) and the inner cylinder chamber (C2) and has flown into the high-pressure
space (S2) is discharged from the discharge pipe (15), is subjected to condensation,
expansion, and evaporation processes in the refrigerant circuit, and then is sucked
into the compressor (1) again.
[0073] Now, it is described how the torque of the electric motor (30) is controlled. In
FIG. 3, it is assumed that FIG. 3(A) corresponds to a rotation angle of 180°, FIG.
3(B) corresponds to a rotation angle of 270°, FIG. 3(C) corresponds to a rotation
angle of 0° (360°), and FIG. 3(D) corresponds to a rotation angle of 90°.
[0074] In the operation described above, the compression torque (i.e., the load torque)
of the drive shaft (33) in one turn of rotation varies as indicated by the solid line
in FIG. 4. Specifically, in the compressor (1) of this embodiment, the compression
torque is highest around a rotation angle of 180°, and is lowest around rotation angles
of 90° and 270°. On the other hand, as indicated by the broken line in FIG. 4, in
a general one-cylinder rotary compressor, the compression torque is highest around
a rotation angle of 180°, and is lowest around a rotation angle of 0° (360°). Comparison
of the torque variation ranges (i.e., a difference between the maximum and minimum
compression torques) in one turn of rotation shows that the torque variation range
(about 1.1 Nm) of the compressor (1) of the present invention is greatly smaller than
the torque variation range (about 2.3 Nm) of the one-cylinder rotary compressor. The
torque variations shown in FIG. 4 are obtained when the operation pressure ratio (i.e.,
the condensation pressure/ the evaporation pressure) occurring in an air conditioner
in an intermediate season is about 1.6.
[0075] Input current of the electric motor (30) is adjusted in accordance with the variation
of the compression torque described above. Specifically, the input current value is
largest when the compression torque is highest, and the input current value is smallest
when the compression torque is lowest. In this manner, in one turn of rotation of
the drive shaft (33), the input current of the electric motor (30) varies from the
minimum value to the maximum value. However, the amount of variation in this input
current (i.e., the degree of control) is smaller than that in the one-cylinder rotary
compressor. That is, the one-cylinder rotary compressor exhibits a wide variation
range of the compression torque in one turn of rotation, and thus the amount of variation
in the input current is also large accordingly.
[0076] In general, as the amount of variation in the input current of an electric motor
decreases, the efficiency in the electric motor increases (i.e., the amount of decrease
in the efficiency decreases). This shows that even with the same torque control, the
efficiency of the electric motor (30) less decreases in the compressor (1) of the
present invention than in the one-cylinder rotary compressor. As a result, the compressor
(1) can operate with energy saving as a whole.
[0077] Now, a relationship among the volume ratio Vr between the outer cylinder chamber
(C1) and the inner cylinder chamber (C2), the torque variation ratio, and the vibration
ratio is described.
[0078] First, FIG. 5 shows a relationship between the volume ratio Vr (Vin/Vout) between
the outer cylinder chamber (C1) and the inner cylinder chamber (C2), and the torque
variation range. In FIG. 5, torque variation ranges are shown for five patterns of
volume ratios Vr (Vin/Vout): 50/50=1, 40/60=0.66, 25/75=0.33, 15/85=0.17, and 0/100=0.
The pattern of the volume ratio Vr (Vin/Vout)=0/100 corresponds to a one-cylinder
rotary compressor. The torque variations shown in FIG. 5 are obtained when the operating
pressure ratio (i.e., the condensation pressure/the evaporation pressure) occurring
in an air conditioner in rated operation is about 3.
[0079] Specifically, the case of the volume ratio Vr=0/100 exhibits the largest torque variation
range, and the case of the volume ratio Vr=50/50 exhibits the smallest torque variation
range. That is, as the volume ratio Vr approaches 1 (one), the torque variation range
decreases. Accordingly, as the volume ratio Vr approaches 1 (one), vibration caused
by a torque variation is more greatly suppressed.
[0080] It is also shown that the period (i.e., an interval between two valleys sandwiching
one peak of a main torque variation) of a main torque variation becomes shorter as
the volume ratio Vr approaches 1 (one). For example, the period ("c" in FIG. 5) of
the main torque variation in the case of the volume ratio Vr=50/50 is shorter than
the period ("b" in FIG. 5) of the main torque variation in the case of the volume
ratio Vr=25/75. This period ("b" in FIG. 5) of the main torque variation is shorter
than the period ("a" in FIG. 5) of the main torque variation in the case of the volume
ratio Vr=0/100. As the period of the main torque variation increases, the motor vibrates
more slowly, and thus the amplitude of this vibration increases. In general, the amplitude
increases in proportion to the square of the period (=1/frequency).
[0081] FIG. 6 shows how the amounts of decrease in a torque variation ratio, a vibration
ratio, and a motor efficiency (i.e., an electric-motor efficiency) are associated
with the volume ratio Vr. In FIG. 6, the torque variation ratio and the vibration
ratio are expressed as ratios of the torque variation range and vibration to the volume
ratio Vr, with the torque variation range and the vibration in the case of the volume
ratio Vr=0/100 being defined as "1". The amount of decrease in the motor efficiency
is obtained when the rotation speed variation is suppressed to the lowest degree with
torque control. In FIG. 6, the amount of decrease in the motor efficiency (i.e., the
electric motor efficiency) is indicated by a solid line, the torque variation ratio
is indicated by a broken line, and the vibration ratio is indicated by a dash-dotted
line.
[0082] Specifically, as the volume ratio Vr approaches 1 (one), the torque variation ratio
and the vibration ratio decrease. The amount of decrease in the motor efficiency is
approximately 0%, i.e., is smallest, when the volume ratio Vr is 1 (one), and increases
as the volume ratio Vr decreases. In addition, it is shown that the motor efficiency
gradually decreases while the volume ratio Vr is in the range from 1.0 to 0.6, and
steeply decreases when the volume ratio Vr is less than 0.6. In this manner, while
the volume ratio Vr is in the range from 0.6 to 1.0, torque control can be performed
with a small amount of decrease in the motor efficiency, and thus vibration can be
suppressed as compared to the one-cylinder rotary compressor.
[0083] As described above, in the compressor (1) of this embodiment, torque control of the
electric motor (30) enables suppression of a decrease in the efficiency of the electric
motor (30) more greatly than torque control of a one-cylinder rotary compressor. In
addition, despite performing torque control with a small amount of decrease in the
efficiency of the electric motor (30), vibration can be further suppressed by setting
the volume ratio Vr between the outer cylinder chamber (C1) and the inner cylinder
chamber (C2) at about 0.7. As a result, vibration of the compressor (1) can be suppressed,
and operation with energy saving can be achieved.
[0084] In this embodiment, a brushless direct-current (DC) motor having a higher efficiency
than an alternating-current (AC) motor is employed as the electric motor (30), and
thus high efficiency can be maintained even during low-speed operation necessary for
an air conditioner incorporating the compressor (1) of the present invention in intermediate
seasons, thus further saving energy.
[0085] In a conventional two-cylinder rotary compressor, two cylinders having a volume ratio
Vr of 1 : 1 are disposed longitudinally, and thus a crank mechanism such as a rotary
piston and an eccentric shaft is needed for each cylinder. On the other hand, in the
compressor (1) of the present invention, one cylinder is partitioned into the outer
cylinder chamber (C1) and the inner cylinder chamber (C2), and these cylinder chambers
share the annular piston (22). As a result, only one crank mechanism is sufficient
for the compressor (1), thus achieving cost reduction.
[0086] In addition, if the cylinder (21) is made of an aluminum alloy, centrifugal force
during rotation is reduced. In this case, vibration in high-speed operation can be
suppressed, and warping of the drive shaft (33) is also suppressed. Accordingly, highly
efficient operation with small vibration can be achieved in a wide range.
«EMBODIMENT 2»
[0087] As illustrated in FIGS. 7 and 8, a second embodiment is obtained by modifying the
configuration of the compression mechanism (20) of the first embodiment. Specifically,
in this embodiment, an annular piston (52) is movable, a cylinder (21) is stationary,
and the annular piston (52) eccentrically rotates relative to the cylinder (21).
[0088] A compression mechanism (20) of the second embodiment includes an upper housing (16)
and a piston assembly (55). The upper housing (16) is continuous to the cylinder (21).
The piston assembly (55) is configured to eccentrically rotate relative to the cylinder
(21). In this embodiment, the lower housing (17) is omitted.
[0089] The cylinder (21) includes outer cylinder portion (24) and an inner cylinder portion
(25). The outer cylinder portion (24) and the inner cylinder portion (25) are annular,
and have the same axis. These outer and inner cylinder portions (24) and (25) are
provided on the lower surface of a head (26) of the upper housing (16). An annular
cylinder chamber (C1, C2) is formed between the inner peripheral surface of the outer
cylinder portion (24) and the outer peripheral surface of the inner cylinder portion
(25).
[0090] The piston assembly (55) includes: a head (51); the annular piston (52) positioned
upright on, and continuous to, the upper surface of the head (51); and a cylindrical
piston (53). The piston assembly (55) is made of cast steel or an aluminum alloy.
The annular piston (52) has an inner diameter greater than the outer diameter of the
cylindrical piston (53), and has the same axis as the cylindrical piston (53). The
piston assembly (55) is configured such that the annular piston (52) is placed in
the annular cylinder chamber (C1, C2) to partition the cylinder chamber (C1, C2) into
an outer cylinder chamber (C1) and an inner cylinder chamber (C2). Specifically, the
head (26) of the upper housing (16) serves as a first block member for blocking an
end of the cylinder chamber (C1, C2). The head (51) of the piston assembly (55) serves
as a second block member for blocking the other end of the cylinder chamber (C1, C2).
The cylindrical piston (53) is located in the inner cylinder portion (25).
[0091] In this embodiment, the volume Vout of the outer cylinder chamber (C1) is also larger
than the volume Vin of the inner cylinder chamber (C2), and the volume ratio Vr (Vin/Vout)
of the inner cylinder chamber (C2) to the outer cylinder chamber (C1) is also set
at about 0.7.
[0092] An eccentric part (33a) is formed at the top of a drive shaft (33) of an electric
motor (30), and is coupled to the piston assembly (55). Specifically, this eccentric
part (33a) of the drive shaft (33) is rotatably fitted in a fitting part (54) which
has a cylindrical shape and is continuous to the lower surface of the piston assembly
(55). In this manner, the piston assembly (55) is caused to eccentrically rotate relative
to the cylinder (21) by the rotation of the drive shaft (33).
[0093] Next, it is described how this compressor (1) operates with reference to FIG. 9.
Advantages of the cylinder chamber (C1, C2) in this operation are substantially the
same as those in the first embodiment.
[0094] Specifically, in the outer cylinder chamber (C1), the volume of a low-pressure chamber
(C1-Lp) is approximately smallest in the state shown in FIG. 9(D). While the drive
shaft (33) rotates to shift from this state to the states of FIGS. 9(A), 9(B), and
9(C) in this order, the volume of the low-pressure chamber (C1-Lp) increases, and
refrigerant is sucked into the low-pressure chamber (C1-Lp). When the drive shaft
(33) takes one turn, the low-pressure chamber (C1-Lp) changes to a high-pressure chamber
(C1-Hp). Then, while the drive shaft (33) further rotates, the volume of the high-pressure
chamber (C1-Hp) decreases, thereby compressing refrigerant.
[0095] On the other hand, in the inner cylinder chamber (C2), the volume of the low-pressure
chamber (C2-Lp) is approximately smallest in the state shown in FIG. 9(B). While the
drive shaft (33) rotates to shift from this state to the states of FIGS. 9(C), 9(D),
and 9(A) in this order, the volume of the low-pressure chamber (C2-Lp) increases,
and thus refrigerant is sucked into the low-pressure chamber (C2-Lp). When drive shaft
(33) takes one turn, the low-pressure chamber (C2-Lp) changes to a high-pressure chamber
(C2-Hp). Then, while the drive shaft (33) further rotates, the volume of the high-pressure
chamber (C2-Hp) decreases, thereby compressing refrigerant.
[0096] As in the first embodiment, torque control of the electric motor (30) is performed
by a controller (50) in this embodiment. Accordingly, as compared to a case where
torque control is performed on a one-cylinder compressor, decrease in the efficiency
of the electric motor (30) is greatly suppressed, thus achieving energy saving of
the compressor (1).
[0097] As in the first embodiment, if the piston assembly (55) is made of an aluminum alloy,
vibration and warping of the drive shaft (33) are suppressed during high-speed operation,
thus performing highly efficient operation with small vibration in a wide range. Other
configurations, operation, and advantages are similar to those in the first embodiment.
«EMBODIMENT 3»
[0098] As illustrated in FIGS. 10 and 11, a third embodiment is obtained by modifying the
compression mechanism (20) of the first embodiment such that the compression mechanism
(20) performs two-stage compression on refrigerant. Specifically, in a compression
mechanism (20) of the third embodiment, an outer cylinder chamber (C1) serves as a
low-stage compression chamber, and an inner cylinder chamber (C2) serves as a high-stage
compression chamber.
[0099] A compressor (1) is used for, for example, a refrigerant circuit using carbon dioxide
(CO
2) as refrigerant and operating in a two-stage compression one-stage expansion cycle.
Although not shown, in this refrigerant circuit, the compressor (1), a heat dissipater
(a gas cooler), a receiver, an intermediate cooler, an expansion valve, and an evaporator
are sequentially connected to each other by refrigerant pipes. In this refrigerant
circuit, high-pressure refrigerant discharged from the inner cylinder chamber (C2)
of the compressor (1) sequentially flows in the heat dissipater, the receiver, the
expansion valve, and the evaporator, and flows into the outer cylinder chamber (C1)
of the compressor (1). On the other hand, intermediate-pressure refrigerant compressed
in the outer cylinder chamber (C1) flows into the intermediate cooler, and part of
liquid refrigerant from the receiver subjected to pressure reduction also flows into
the intermediate cooler. In this intermediate cooler, the intermediate-pressure refrigerant
from the outer cylinder chamber (C1) is cooled. This cooled intermediate-pressure
refrigerant returns to the inner cylinder chamber (C2), and is compressed again. This
circulation is repeated, thereby cooling the inside air in the evaporator, for example.
[0100] A body (11) of a casing (10) of the compressor (1) is provided with a suction pipe
(14), an inflow pipe (1a), and an outflow pipe (1b). These pipes penetrate the body
(11). The suction pipe (14) is connected to the evaporator, and the inflow pipe (1a)
and the outflow pipe (1b) are connected to the intermediate cooler. A discharge pipe
(15) is provided through an upper head (12) of the casing (10). This discharge pipe
(15) is connected to the heat dissipater.
[0101] An upper housing (16) of the compression mechanism (20) is provided with a cover
plate (18). In the casing (10), space above the cover plate (18) is defined as high-pressure
space (4a), and space below a lower housing (17) is defined as intermediate-pressure
space (4b). An end of the discharge pipe (15) is open to the high-pressure space (4a),
and an end of the outflow pipe (1b) is open to the intermediate-pressure space (4b).
[0102] An intermediate-pressure chamber (4c) and a high-pressure chamber (4d) are formed
between the upper housing (16) and the cover plate (18). The upper housing (16) has
an intermediate-pressure passageway (4e). A pocket (4f) is formed on the outer periphery
of an outer cylinder (24) between the upper housing (16) and the lower housing (17).
The inflow pipe (1a) is connected to an end of the intermediate-pressure passageway
(4e). The suction pipe (14) is connected to the pocket (4f) such that the pocket (4f)
is in a low-pressure atmosphere at a suction pressure.
[0103] The outer cylinder (24) has a first inlet (41a) radially penetrating the outer cylinder
(24). The first inlet (41a) is located at the right of a blade (23) in FIG. 11. That
is, the first inlet (41a) establishes communication among the outer cylinder chamber
(C1), the pocket (4f), and the suction pipe (14).
[0104] The other end of the intermediate-pressure passageway (4e) is configured as a second
inlet (41b). This second inlet (41b) is located at the right of the blade (23), is
open to the inner cylinder chamber (C2), and establishes communication between the
inner cylinder chamber (C2) and the intermediate-pressure space (4b).
[0105] The upper housing (16) has a first discharge port (45a) and a second discharge port
(45b). These discharge ports (45a, 45b) axially penetrate the upper housing (16).
An end of the first discharge port (45a) is open to the high-pressure side of the
outer cylinder chamber (C1), and the other end of the first discharge port (45a) is
open to the intermediate-pressure chamber (4c). An end of the second discharge port
(45b) is open to the high-pressure side of the inner cylinder chamber (C2), and the
other end of the second discharge port (45b) is open to the high-pressure chamber
(4d). Outer ends of the first discharge port (45a) and the second discharge port (45b)
are provided with valves (46) which are reed valves.
[0106] The intermediate-pressure chamber (4c) and the intermediate-pressure space (4b) communicate
with each other with a communication passageway (4g) formed in the upper housing (16)
and the lower housing (17). Although not shown, the high-pressure chamber (4d) communicates
with the high-pressure space (4a) via a high-pressure passageway formed in the cover
plate (18).
[0107] In this embodiment, the volume Vout of the outer cylinder chamber (C1) is also larger
than the volume Vin of the inner cylinder chamber (C2), and the volume ratio Vr (Vin/Vout)
of the inner cylinder chamber (C2) to the outer cylinder chamber (C1) is also set
at about 0.7.
[0108] In this compressor (1), when the electric motor (30) is started, the outer cylinder
(24) and the inner cylinder (25) revolve, while swinging relative to the annular piston
(22), as in the first embodiment. Then, the compression mechanism (20) performs given
compression operation.
[0109] When the volume of the outer cylinder chamber (C1) increases with rotation of a drive
shaft (33), low-pressure refrigerant is sucked into the outer cylinder chamber (C1)
from the suction pipe (14) through the pocket (4f) and the first inlet (41a). Then,
the drive shaft (33) further rotates to cause the volume of the outer cylinder chamber
(C1) to decrease, thereby compressing refrigerant. When the pressure of this outer
cylinder chamber (C1) reaches a given intermediate pressure, and the differential
pressure between the outer cylinder chamber (C1) and the intermediate-pressure chamber
(4c) reaches a set value, the valves (46) are opened. Accordingly, intermediate-pressure
refrigerant is discharged from the outer cylinder chamber (C1) into the intermediate-pressure
chamber (4c), and flows from the outflow pipe (1b) through the intermediate-pressure
space (4b).
[0110] On the other hand, when the volume of the inner cylinder chamber (C2) is caused to
increase by the rotation of the drive shaft (33), intermediate-pressure refrigerant
is sucked into the inner cylinder chamber (C2) from the inflow pipe (1a) through the
intermediate-pressure passageway (4e) and the second inlet (41b). Then, the drive
shaft (33) further rotates to cause the volume of the inner cylinder chamber (C2)
to decrease, thereby compressing refrigerant. When the pressure of this inner cylinder
chamber (C2) reaches a given high pressure, and the differential pressure between
the inner cylinder chamber (C2) and the high-pressure chamber (4d) reaches a set value,
the valves (46) are opened. Accordingly, high-pressure refrigerant is discharged from
the inner cylinder chamber (C2) into the high-pressure chamber (4d), and flows from
the discharge pipe (15) through the high-pressure space (4a). In this manner, in the
compressor (1) of this embodiment, refrigerant compressed in the outer cylinder chamber
(C1) is further compressed in the inner cylinder chamber (C2), thereby performing
two-stage compression.
[0111] In general, an air conditioner (which is herein an inverter air conditioner) is most
frequently operated in a range of low pressure ratios where the operating pressure
ratio is in the range from about 1.6 to about 2.0. The operating pressure ratio herein
is a ratio of a condensation pressure to an evaporation pressure in a refrigerant
circuit.
[0112] As shown in FIG. 12, in a low operating pressure ratio range with a high operating
frequency, as the volume ratio Vr of the inner cylinder chamber (C2) to the outer
cylinder chamber (C1) increases, the compression efficiency increases. For example,
in the case of the volume ratio Vr=0.8, the compression efficiency is highest around
an operating pressure ratio of 1.5, and in the case of the volume ratio Vr=0.6, the
compression efficiency is highest around an operating pressure ratio of 1.9. On the
other hand, in the case of the volume ratio Vr=0.5, the compression efficiency is
highest around an operating pressure ratio of 2.5, and while the operating pressure
ratio decreases to about 2.0 or less, the compression efficiency greatly decreases.
That is, in the two-stage compression using the outer cylinder chamber (C1) and the
inner cylinder chamber (C2), as the volume ratio Vr decreases, the compression ratio
of the entire compression mechanism (20) increases. Then, under operating conditions
with a low operating pressure ratio, overcompression losses are likely to occur, and
thus the compression efficiency is also likely to decrease.
[0113] FIG. 13 shows a relationship between the volume ratio Vr (Vin/Vout) and the torque
variation range: FIG. 14 shows a relationship between the volume ratio Vr (Vin/Vout)
and the torque variation ratio. These graphs show that the torque variation range
and the torque variation ratio in the cases of the volume ratio Vr=0.6 and 0.8 are
smaller than those in the case of the volume ratio Vr=0.5 and 1. Accordingly, such
small torque variations enable suppression of vibration. The torque variation ratio
is the torque variation range expressed in terms of ratio for each volume ratio Vr,
with the torque variation range in the case of the volume ratio Vr=1 set at "1". FIGS.
13 and 14 are obtained by measurement at an operating pressure ratio of about 2.
[0114] In the case of the volume ratio Vr=1, the outer cylinder chamber (C1) and the inner
cylinder chamber (C2) have the same volume. In this case, refrigerant is not compressed
in the outer cylinder chamber (C1), but is compressed only in the inner cylinder chamber
(C2), i.e., is subjected to not two-stage compression but single-stage compression.
That is, the case of the volume ratio Vr=1 substantially corresponds to the case where
single-stage compression is performed by a conventional one-cylinder rotary compressor.
Here, in the case of the volume ratio Vr=0.5, the torque variation range and the torque
variation ratio are greater than those in the case of Vr=1. Specifically, large vibration
requires suppression of the vibration by means of torque control, resulting in that
the torque control might cause the operational efficiency to be lower than that in
the conventional one-cylinder rotary compressor.
[0115] As described above, in the two-stage compression configuration of this embodiment,
the volume ratio Vr between the inner and outer cylinder chambers (C1, C2) are set
in the range from about 0.6 to about 0.8. Accordingly, the compression efficiency
can be enhanced and vibration can be suppressed, as compared to a conventional one-cylinder
compressor.
«EMBODIMENT 4»
[0116] As illustrated in FIGS. 15 and 16, a fourth embodiment is obtained by modifying the
two-stage compression configuration of the compression mechanism (20) of the third
embodiment. Specifically, in the compression mechanism (20) of the third embodiment,
two cylinder chambers (C1, C2) are formed in the same plane. On the other hand, in
the fourth embodiment, a compression mechanism (80) is formed by longitudinally stacking
two cylinder chambers (82a, 82b), and constitutes a so-called two-stage rotary compressor.
[0117] More specifically, a compressor (60) of the fourth embodiment is configured such
that the compression mechanism (80) having a low-stage compression mechanism (80a)
and a high-stage compression mechanism (80b) and an electric motor (65) are housed
in a casing (61) which is a sealed vessel in an elongated cylindrical shape. In the
casing (61), the electric motor (65) is located above the compression mechanism (80).
[0118] A suction pipe (62) penetrates the body of the casing (61), and a discharge pipe
(63) penetrates the body at a position above the suction pipe (62). The discharge
pipe (63) is bent at the inlet side thereof in the casing (61), then extends horizontally,
and is open at the end.
[0119] The electric motor (65) includes a stator (66) and a rotor (67). The stator (66)
is fixed to the inner peripheral surface of the casing (61). The rotor (67) is located
at the inner side of the stator (66). A center portion of the rotor (67) is coupled
to a main shaft portion (71) of a drive shaft (70) which longitudinally extends.
[0120] The drive shaft (70) constitutes a drive shaft. The drive shaft (70) has a first
eccentric part (72) and a second eccentric part (73) which are located in this order
from the bottom. Each of the first eccentric part (72) and the second eccentric part
(73) has a diameter greater than that of the main shaft portion (71), and is eccentric
to the axis of the main shaft portion (71). The first eccentric part (72) and the
second eccentric part (73) are respectively eccentric to the axis of the main shaft
portion (71) in opposite directions. The height of the first eccentric part (72) is
greater than that of the second eccentric part (73).
[0121] The compression mechanism (80) is configured such that a rear head (84), a first
cylinder (81a), a middle plate (86), a second cylinder (81b), and a front head (83)
are stacked in this order. The first cylinder (81 a) houses a first rotary piston
(87a). The second cylinder (81b) houses a second rotary piston (87b).
[0122] The first cylinder (81a), the first rotary piston (87a), the rear head (84), and
the middle plate (86) constitute the low-stage compression mechanism (80a). The second
cylinder (81b), the second rotary piston (87b), the front head (83), and the middle
plate (86) constitute the high-stage compression mechanism (80b). Each of the low-stage
compression mechanism (80a) and the high-stage compression mechanism (80b) is configured
by a rotary fluid machine of a swinging piston type, which is a type of a positive-displacement
fluid machine.
[0123] As illustrated in FIG. 16, the first rotary piston (87a) of the low-stage compression
mechanism (80a) has an annular shape. The first eccentric part (72) is rotatably fitted
in this first rotary piston (87a) of the low-stage compression mechanism (80a). The
second rotary piston (87b) of the high-stage compression mechanism (80b) also has
an annular shape. The second eccentric part (73) is rotatably fitted in this second
rotary piston (87b) of the high-stage compression mechanism (80b).
[0124] The inner peripheral surfaces of the rotary pistons (87a, 87b) are respectively in
slidable contact with the outer peripheral surfaces of the eccentric parts (72, 73),
and the outer peripheral surfaces of the rotary pistons (87a, 87b) are respectively
in slidable contact with the inner peripheral surfaces of the cylinders (81a, 81b).
A cylinder chamber (82a, 82b) is formed between the outer peripheral surface of the
rotary piston (87a, 87b) and the inner peripheral surface of the cylinder (81a, 81b).
A blade (74) in the shape of a flat plate projects from a side surface of each of
the rotary pistons (87a, 87b). The blade (74) is supported by the cylinder (81a, 81b)
with a swing bushing (75) interposed therebetween. The swing bushing (75) is provided
between a discharge port (89a, 89b) and a suction port (88a, 88b) which will be described
later. The blade (74) partitions the cylinder chamber (82a, 82b) into a high-pressure
side and a low-pressure side.
[0125] In this manner, the compression mechanism (80) is configured such that rotation of
the eccentric part (72, 73) causes the rotary piston (87a, 87b) to revolve and swing
in the cylinder chamber (82a, 82b). The rotational phases of the rotary pistons (87a,
87b) shift from each other by 180°.
[0126] The first cylinder (81a) of the low-stage compression mechanism (80a) and the second
cylinder (81 b) of the high-stage compression mechanism (80b) have the same inner
diameter. The first rotary piston (87a) of the low-stage compression mechanism (80a)
and the second rotary piston (87b) of the high-stage compression mechanism (80b) have
the same outer diameter. The height of the first cylinder (81 a) of the low-stage
compression mechanism (80a) is greater than that of the second cylinder (81b) of the
high-stage compression mechanism (80b).
[0127] The middle plate (86) has an annular intermediate passageway (90). A discharge port
(89a) of the low-stage compression mechanism (80a) is formed in the middle plate (86).
This discharge port (89a) allows the high-pressure side of the first cylinder chamber
(82a) of the low-stage compression mechanism (80a) to communicate with the intermediate
passageway (90). On the other hand, a discharge port (89b) of the high-stage compression
mechanism (80b) is formed in the front head (83). This discharge port (89b) allows
the high-pressure side of the second cylinder chamber (82b) of the high-stage compression
mechanism (80b) to communicate with the space in the casing (61). These discharge
ports (89a, 89b) have discharge valves (not shown) for opening and closing the respective
outlets of the discharge ports (89a, 89b).
[0128] The first cylinder (81a) of the low-stage compression mechanism (80a) has a suction
port (88a). This suction port (88a) radially penetrates the first cylinder (81a).
The terminal end of the suction port (88a) is open to the first cylinder chamber (82a).
This suction port (88a) is connected to the suction pipe (62). The second cylinder
(81b) of the high-stage compression mechanism (80b) has a suction port (88b) extending
from the middle plate (86). The starting end of the suction port (88b) is open to
the intermediate passageway (90), and the terminal end of the suction port (88b) is
open to the second cylinder chamber (82b).
[0129] An oil sump for storing lubricating oil is formed at the bottom of the casing (61).
A centrifugal oil supply pump (92) immersed in the oil sump is provided at the bottom
of the drive shaft (70). This oil supply pump (92) longitudinally extends in the drive
shaft (70), and is connected to an oil supply passageway (91) which communicates with
the low-stage compression mechanism (80a) and the high-stage compression mechanism
(80b). The oil supply pump (92) is configured to supply lubricating oil in the oil
sump to a sliding portion of the low-stage compression mechanism (80a) and to a sliding
portion of the high-stage compression mechanism (80b) through the oil supply passageway
(91).
[0130] In this embodiment, the volume V1 of the first cylinder chamber (82a) is also larger
than the volume V2 of the second cylinder chamber (82b), and the volume ratio Vr (V2/V1)
of the second cylinder chamber (82b) to the first cylinder chamber (82a) is also set
at about 0.7.
[0131] In this compressor (60), when the electric motor (65) is started, the rotary pistons
(87a, 87b) revolve, while swinging in the cylinder chambers (81a, 81b). Then, the
compression mechanism (80) performs given compression operation.
[0132] This compression operation is described with reference to FIG. 16. In FIG. 16, the
rotary piston (87a, 87b) swings, while rotating in the clockwise direction. The position
in which the rotary piston (87a, 87b) is in contact with the top dead center is defmed
as a rotation angle of 0°, and the position in which the rotary piston (87a, 87b)
is in contact with the bottom dead center is defined as a rotation angle of 180°.
First, when the drive shaft (70) slightly rotates from the state of a rotation angle
of 0° so that the contact point between the first rotary piston (87a) and the first
cylinder (81a) passes by the opening of the suction port (88a), refrigerant starts
to flow from the suction port (88a) to the first cylinder chamber (82a). The refrigerant
continues to flow into the first cylinder chamber (82a) until the rotation angle of
the drive shaft (70) reaches 360°.
[0133] Subsequently, in a state where the flow of refrigerant into the first cylinder chamber
(82a) is finished (i.e., the rotation angle of the drive shaft (70) is 360°), when
the drive shaft (70) slightly rotates so that the contact point between the first
rotary piston (87a) and the first cylinder (81a) passes by the opening of the suction
port (88a), confinement of refrigerant in the first cylinder chamber (82a) is completed.
Then, from this state, the drive shaft (70) further rotates, thereby starting compression
of refrigerant. When the pressure of refrigerant in the first cylinder chamber (82a)
exceeds the pressure of refrigerant in the intermediate passageway (90), the discharge
valve is opened, thereby discharging intermediate-pressure refrigerant from the discharge
port (89a) into the intermediate passageway (90). The discharge of refrigerant continues
until the rotation angle of the drive shaft (70) reaches 360°.
[0134] On the other hand, in the high-stage compression mechanism (80b), the rotation of
the drive shaft (70) causes intermediate-pressure refrigerant in the intermediate
passageway (90) to flow from the suction port (88b) into the second cylinder chamber
(82b). Specifically, when the contact point between the second rotary piston (87b)
and the second cylinder (81b) passes by the opening of the suction port (88b), refrigerant
starts to flow from the intermediate passageway (90) into the second cylinder chamber
(82b). The flow of intermediate-pressure refrigerant continues until the rotation
angle of the drive shaft (70) reaches 360°.
[0135] Thereafter, when the contact point between the second rotary piston (87b) and the
second cylinder (81b) passes by the opening of the suction port (88b) so that confinement
of refrigerant in the second cylinder chamber (82b) is completed, compression of refrigerant
starts. Then, when the pressure of refrigerant in the second cylinder chamber (82b)
exceeds the pressure of refrigerant in the space in the casing (61), the discharge
valve is opened, thereby discharging the high-pressure refrigerant from the discharge
port (89b) into the space in the casing (61). The discharge of refrigerant continues
until the rotation angle of the drive shaft (70) reaches 360°. The refrigerant discharged
into the space in the casing (61) is discharged from the discharge pipe (63) to the
refrigerant circuit. In this manner, in the compressor (60) of this embodiment, refrigerant
compressed in the low-stage first cylinder chamber (82a) is further compressed in
the high-stage second cylinder chamber (82b), thereby performing two-stage compression.
[0136] As in the third embodiment, the relationship between the volume ratio Vr (V2n/V1)
and the torque variation range and the relationship between the volume ratio Vr (V2n/V1)
and the torque variation ratio also apply to the relationships shown in FIGS. 13 and
14. That is, the torque variation range and the torque variation ratio in the case
of the volume ratio Vr=0.6 and 0.8 are smaller than those in the case of the volume
ratio Vr= 0.5 and 1. Accordingly, such small torque variations enable suppression
of vibration. Thus, in the two-stage compression configuration in which the low-stage
and high-stage cylinder chambers (82a, 82b) are longitudinally stacked, as long as
the volume ratio Vr between these cylinder chambers (82a, 82b) are set in the range
from about 0.6 to about 0.8, the compression efficiency can be enhanced and vibration
can be suppressed, as compared to a conventional one-cylinder compressor.
INDUSTRIAL APPLICABILITY
[0137] As described above, the present invention is useful for rotary compressors each including
two cylinder chambers whose volume varies with eccentric rotation of a piston.