U.S. patent application number 17/633907 was filed with the patent office on 2022-09-15 for refrigerating apparatus using non-azeotropic mixed refrigerant.
The applicant listed for this patent is LG ELECTRONICS INC.. Invention is credited to Kyeongyun KIM, Yongjoo PARK, Changho SEO, Minho SONG.
Application Number | 20220290904 17/633907 |
Document ID | / |
Family ID | 1000006433032 |
Filed Date | 2022-09-15 |
United States Patent
Application |
20220290904 |
Kind Code |
A1 |
KIM; Kyeongyun ; et
al. |
September 15, 2022 |
REFRIGERATING APPARATUS USING NON-AZEOTROPIC MIXED REFRIGERANT
Abstract
A refrigerating apparatus using a non-azeotropic mixed
refrigerant may include a compressor operable in a continuous
operation mode and configured to compress the non-azeotropic mixed
refrigerant, a condenser configured to condense the refrigerant
compressed by the compressor, an expander configured to expand the
refrigerant condensed by the condenser, and an evaporator
configured to evaporate the refrigerant expanded by the expander. A
pressure difference (.DELTA.P) of the non-azeotropic mixed
refrigerant has a value included in a range of 340
kPa<.DELTA.P<624.7 kPa. Therefore, reliability of components,
such as a piston, in the refrigerating apparatus using the
non-azeotropic mixed refrigerant may be further improved.
Inventors: |
KIM; Kyeongyun; (Seoul,
KR) ; SEO; Changho; (Seoul, KR) ; PARK;
Yongjoo; (Seoul, KR) ; SONG; Minho; (Seoul,
KR) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
LG ELECTRONICS INC. |
Seoul |
|
KR |
|
|
Family ID: |
1000006433032 |
Appl. No.: |
17/633907 |
Filed: |
August 20, 2020 |
PCT Filed: |
August 20, 2020 |
PCT NO: |
PCT/KR2020/011141 |
371 Date: |
February 8, 2022 |
Current U.S.
Class: |
1/1 |
Current CPC
Class: |
F25B 5/04 20130101; F25B
31/002 20130101; F25B 31/023 20130101; F25B 9/006 20130101; F04B
39/121 20130101; F25B 2400/073 20130101 |
International
Class: |
F25B 31/02 20060101
F25B031/02; F25B 31/00 20060101 F25B031/00; F25B 9/00 20060101
F25B009/00; F25B 5/04 20060101 F25B005/04; F04B 39/12 20060101
F04B039/12 |
Foreign Application Data
Date |
Code |
Application Number |
Aug 21, 2019 |
KR |
10-2019-0102338 |
Claims
1. A refrigerating apparatus including a non-azeotropic mixed
refrigerant, the refrigerating apparatus comprising: a compressor
operable in a continuous operation mode and to compress the
non-azeotropic mixed refrigerant; a condenser to condense the
non-azeotropic mixed refrigerant compressed by the compressor; an
expander to expand the non-azeotropic mixed refrigerant condensed
by the condenser; and an evaporator to evaporate the non-azeotropic
mixed refrigerant expanded by the expander, wherein a pressure
difference (.DELTA.P) of the non-azeotropic mixed refrigerant is in
a range of 340 kPa<.DELTA.P<624.7 kPa.
2. The refrigerating apparatus according to claim 1, wherein a
condensing pressure (Pd) of the non-azeotropic mixed refrigerant is
in a range of 393.4 kPa<Pd<745.3 kPa.
3. The refrigerating apparatus according to claim 1, wherein an
evaporation pressure (Ps) of the non-azeotropic mixed refrigerant
is in a range of 53.5 kPa<Ps<120.5 kPa.
4. The refrigerating apparatus according to claim 1, wherein, in
the continuous operation mode, the compressor is controlled to
operate even when an interior temperature of the refrigerating
apparatus is in a target temperature range.
5. The refrigerating apparatus according to claim 1, wherein the
compressor comprises a linear compressor.
6. The refrigerating apparatus according to claim 5, wherein the
linear compressor comprises: a shell including a suction inlet; a
cylinder disposed in the shell to define a refrigerant compression
space; a frame coupled to an outer side of the cylinder; a piston
to reciprocate within the cylinder; a discharge valve movably
coupled to the cylinder to selectively discharge the non-azeotropic
mixed refrigerant compressed in the refrigerant compression space;
and a passage that extends into a space between the cylinder and
the frame and through which at least a portion of the
non-azeotropic mixed refrigerant discharged from the discharge
valve flows.
7. The refrigerating apparatus according to claim 6, wherein the
cylinder comprises: a cylinder body including at least one nozzle;
and a cylinder flange that extends radially outward from the
cylinder body.
8. The refrigerating apparatus according to claim 7, wherein the
frame comprises: a frame body that surrounds the cylinder body; a
recessed portion which receives the cylinder flange; and a seating
portion that faces a seating surface of the cylinder flange.
9. The refrigerating apparatus according to claim 8, wherein the
passage comprises a first passage between an outer circumferential
surface of the cylinder flange and an inner circumferential surface
of the recessed portion.
10. The refrigerating apparatus according to claim 9, wherein the
passage comprises a second passage between the seating surface of
the cylinder flange and a seating surface of the frame.
11. The refrigerating apparatus according to claim 10, wherein the
passage comprises a third passage that extends into a space between
an outer circumferential surface of the cylinder body and an inner
circumferential surface of the frame body.
12. The refrigerating apparatus according to claim 5, wherein the
linear compressor comprises: a cylinder including a cylinder
stepped portion on an inner circumferential surface thereof; a
piston to reciprocate within the cylinder and including a piston
stepped portion on an outer circumferential surface thereof,
wherein a low pressure is formed between the piston stepped portion
and the cylinder stepped portion when the piston moves in a first
direction and a high pressure is formed between the piston stepped
portion and the cylinder stepped portion when the piston moves in a
second direction; an oil suction passage to allow oil to flow
between the cylinder stepped portion and the piston stepped
portion; and an oil discharge passage to allow the oil between the
cylinder stepped portion and the piston stepped portion to be
discharged outside of the cylinder.
13. The refrigerating apparatus according to claim 1, wherein the
non-azeotropic mixed refrigerant comprises a first hydrocarbon that
is isobutane and a second hydrocarbon that is propane.
14. The refrigerating apparatus according to claim 13, wherein the
isobutane has a weight ratio of
76%.ltoreq.isobutane.ltoreq.87%.
15. A refrigerating apparatus including a non-azeotropic mixed
refrigerant, the refrigerating apparatus comprising: a linear
compressor to compress the non-azeotropic mixed refrigerant; a
condenser to condense the non-azeotropic mixed refrigerant
compressed by the compressor; an expander to expand the
non-azeotropic mixed refrigerant condensed by the condenser; and an
evaporator to evaporate the non-azeotropic mixed refrigerant
expanded by the expander, wherein a pressure difference (.DELTA.P)
of the non-azeotropic mixed refrigerant is in a range of 340
kPa<.DELTA.P<624.7 kPa.
16. The refrigerating apparatus according to claim 15, wherein the
linear compressor comprises: a piston that reciprocates; and a
cylinder to guide the piston, wherein the non-azeotropic mixed
refrigerant of high pressure compressed by the piston is guided to
an inner surface of the cylinder to cause an outer surface of the
piston to be spaced from the inner surface of the cylinder.
17. The refrigerating apparatus according to claim 15, wherein the
non-azeotropic mixed refrigerant comprises at least two
hydrocarbons, the at least two hydrocarbons comprising: at least
one first hydrocarbon having an evaporation temperature of
-12.degree. C. or more at 1 bar; and at least one second
hydrocarbon having an evaporation temperature of -50.degree. C. or
more and less than -12.degree. C. at 1 bar, and wherein a gliding
temperature difference is 4.degree. C. or more.
18. The refrigerating apparatus according to claim 17, wherein a
weight ratio of the at least one first hydrocarbon is 50% or
more.
19. The refrigerating apparatus according to claim 15, wherein the
linear compressor comprises: a piston that reciprocates to compress
the non-azeotropic mixed refrigerant; and a cylinder to guide the
piston, wherein oil pumped by a pressure difference of the
non-azeotropic mixed refrigerant is present at contact surfaces
between the piston and the cylinder.
20. A refrigerating apparatus including a non-azeotropic mixed
refrigerant, the refrigerating apparatus comprising: a compressor
to compress the non-azeotropic mixed refrigerant; a condenser to
condense the non-azeotropic mixed refrigerant compressed by the
compressor; an expander to expand the non-azeotropic mixed
refrigerant condensed by the condenser; and an evaporator to
evaporate the non-azeotropic mixed refrigerant expanded by the
expander and to provide cold air to an inner space of the
refrigerating apparatus, wherein the compressor is selectively
operable in an intermittent operation mode and a continuous
operation mode, and wherein in the continuous operation mode, the
compressor is controlled to operate continuously even when a
temperature of the inner space of the refrigerating apparatus is
within a target temperature range.
Description
TECHNICAL FIELD
[0001] A refrigerating apparatus using a non-azeotropic mixed
refrigerant is disclosed herein.
BACKGROUND ART
[0002] A refrigerating apparatus has a cavity an inner space of
which is maintained at a low temperature. In the refrigerating
apparatus, a refrigeration cycle is provided so as to maintain the
cavity at a low temperature. In the refrigeration cycle, a
refrigerant circulates through processes of compression,
condensation, expansion, and evaporation.
[0003] There are various types of refrigerants. A mixed refrigerant
is a refrigerant in which two or more types of refrigerants are
mixed. Mixed refrigerants include azeotropic mixed refrigerant and
non-azeotropic mixed refrigerant.
[0004] The azeotropic mixed refrigerant is a refrigerant that
changes phase without changing a composition of a gas phase and a
liquid phase, similar to a single refrigerant. An evaporation
temperature of the azeotropic mixed refrigerant is constant between
an inlet and an outlet of the evaporator.
[0005] In the non-azeotropic mixed refrigerant, a refrigerant
having a low boiling point evaporates first, and a refrigerant
having a high boiling point evaporates later. Therefore, the
non-azeotropic mixed refrigerant has different gas phase and liquid
phase compositions during evaporation, and the evaporation
temperature is low at the inlet of the evaporator and high at the
outlet of the evaporator.
[0006] The non-azeotropic mixed refrigerant has a gliding
temperature difference (GTD), which is a characteristic in which
the temperature changes at equal pressure during phase change. When
the non-azeotropic mixed refrigerant is used, the temperature rises
when evaporation occurs at equal pressure, and conversely, the
temperature decreases during condensation at equal pressure. In
other words, the gliding temperature difference of the refrigerant
occurs when the state changes from a saturated liquid to a
saturated gas.
[0007] A thermal efficiency of a heat exchanger may be improved
using this phenomenon. For example, the non-azeotropic mixed
refrigerant may form a Lorentz cycle in which a temperature between
a refrigerant and a heat source is balanced, and efficiency may be
improved by reducing irreversible heat exchange.
[0008] As an existing technique for applying the non-azeotropic
mixed refrigerant, applicant proposed a capillary structure of a
refrigerating apparatus in Korean Patent Registration No. 0119839,
which is hereby incorporated by reference.
DISCLOSURE
Technical Problem
[0009] Although thermal efficiency may be improved using a
non-azeotropic mixed refrigerant, an optimum composition and
optimum operating conditions of the non-azeotropic mixed
refrigerant suitable for a refrigeration cycle applied to a
refrigerating apparatus are unknown.
Technical Solution
[0010] According to embodiments disclosed herein, a refrigerating
apparatus using a non-azeotropic mixed refrigerant may include a
compressor operable in a continuous operation mode and configured
to compress a non-azeotropic mixed refrigerant, a condenser
configured to condense the refrigerant compressed by the
compressor, an expander configured to expand the refrigerant
condensed by the condenser, and an evaporator configured to
evaporate the refrigerant expanded by the expander. A pressure
difference (.DELTA.P) of the non-azeotropic mixed refrigerant may
have a value included in a range of 340 kPa<.beta.P<624.7
kPa. Therefore, it is possible to reduce friction occurring during
operation of the piston in the compressor. This operation mode may
obtain a greater advantage when the compressor is operated in the
continuous operation mode during operation of the refrigerating
apparatus. The continuous operation mode is an operation mode
corresponding to an intermittent operation mode and may indicate a
state in which the compressor is continuously operated without
turning off the compressor even when a current high interior
temperature is within a target temperature range.
[0011] A condensing pressure (Pd) of the non-azeotropic mixed
refrigerant may have a value included in a range of 393.4
kPa<Pd<745.3 kPa. Thus, the condensing pressure realized by
the non-azeotropic mixed refrigerant may be suitably used for the
compressor.
[0012] An evaporation pressure (Ps) of the non-azeotropic mixed
refrigerant may have a value included in a range of 53.5
kPa<Ps<120.5 kPa. Thus, the evaporative pressure realized by
the non-azeotropic mixed refrigerant may be suitably used for the
compressor.
[0013] In the continuous operation mode, the compressor may be
operated even when an interior temperature is in a satisfaction
temperature range. In the continuous operation mode, due to the
high pressure difference, oil circulation and a piston lift
pressure of a gas bearing may be more reliably obtained.
[0014] When the compressor is a linear compressor, it is possible
to reliably perform a frictional force reduction operation in
operation of the linear compressor by further utilizing action due
to the pressure difference of the non-azeotropic mixed refrigerant.
The linear compressor may include a shell provided with a suction
portion or inlet, a cylinder provided in the shell to define a
refrigerant compression space, a frame coupled to an outer side of
the cylinder, a piston provided to be able to reciprocate within
the cylinder in an axial direction, a discharge valve movably
coupled to the cylinder to selectively discharge a refrigerant
compressed in the refrigerant compression space, and a passage that
extends into a space between the cylinder and the frame and through
which at least a part or portion of the refrigerant discharged from
the discharge valve may flow. Lubrication of the piston may be
performed more smoothly.
[0015] The cylinder may include a cylinder body in which a nozzle
portion or nozzle may be formed, and a cylinder flange portion or
flange that extends outward from the cylinder body. The frame may
include a frame body that surrounds the cylinder body, a recessed
portion or recess into which the cylinder flange portion is
inserted, and a seating portion or seat that faces a seating
surface of the cylinder flange portion. Therefore, an internal
configuration of the linear compressor may be firmly supported.
[0016] The passage may include a first passage formed between an
outer circumferential surface of the cylinder flange portion and an
inner circumferential surface of the recessed portion. Thus, it is
possible to provide a passage through which the compressed high
pressure non-azeotropic mixed refrigerant may be bypassed.
[0017] The passage may include a second passage formed between the
seating surface of the cylinder flange portion and a seating
surface of the frame. Thus, the high pressure non-azeotropic mixed
refrigerant passing through the first passage may be guided.
[0018] The passage may include a third passage that extends into a
space between an outer circumferential surface of the cylinder body
and an inner circumferential surface of the frame body. The high
pressure non-azeotropic mixed refrigerant may be guided to a
plurality of locations in a longitudinal direction of the piston
and cylinder.
[0019] The linear compressor may be provided with an assembly
tolerance. In this case, there is no difficulty in passing the
gas-phase non-azeotropic mixed refrigerant.
[0020] The linear compressor may include a cylinder provided with a
cylinder stepped portion on an inner circumferential surface, a
piston arranged to be able to reciprocate with the cylinder and
provided with a piston stepped portion on an outer circumferential
surface, the piston stepped portion forming a low pressure between
the piston stepped portion and the cylinder stepped portion when
moving in one or a first direction and forming a high pressure
between the piston stepped portion and the cylinder stepped portion
when moving in the other or a second direction, an oil suction
passage formed to allow oil to flow between the cylinder stepped
portion and the piston stepped portion, and an oil discharge
passage formed to allow oil between the cylinder stepped portion
and the piston stepped portion to be discharged outside of the
cylinder. In a case of the linear compressor, oil circulation may
be more smoothly performed by a high pressure difference of the
non-azeotropic mixed refrigerant. Therefore, reliability of the
refrigerating apparatus using the non-azeotropic mixed refrigerant
may be improved.
[0021] The non-azeotropic mixed refrigerant may include a first
hydrocarbon and a second hydrocarbon. The first hydrocarbon may be
isobutane, and the second hydrocarbon may be propane. As an optimum
gliding temperature difference may be obtained, a high-efficiency
refrigerating system may be obtained.
[0022] The isobutane may be provided in a weight ratio of 76%
isobutane 87%. Minimum compression work of the refrigeration cycle,
compatibility of a production facility of a refrigerating system, a
low purchase cost of refrigerant, a high safety of the
refrigerating system, efficiency of the refrigeration cycle, and
convenience of handling the refrigerant may be obtained.
[0023] According to embodiments disclosed herein, a refrigerating
apparatus using a non-azeotropic mixed refrigerant may include a
linear compressor configured to compress a non-azeotropic mixed
refrigerant, a condenser configured to condense the non-azeotropic
mixed refrigerant compressed by the compressor, an expander
configured to expand the non-azeotropic mixed refrigerant condensed
by the condenser, and an evaporator configured to evaporate the
non-azeotropic mixed refrigerant expanded by the expander. A
pressure difference (.DELTA.P) of the non-azeotropic mixed
refrigerant may be 340 kPa<.beta.P<624.7 kPa. It is possible
to reliably perform a friction reducing action between the piston
and the cylinder in the linear compressor.
[0024] The linear compressor may include a piston configured to
reciprocate, and a cylinder configured to guide the piston. The
non-azeotropic mixed refrigerant of high pressure compressed by the
piston may be guided to an inner surface of the cylinder to cause
an outer surface of the piston to be lifted on the inner surface of
the cylinder. In this case, operational reliability of the linear
compressor may be increased by a lift pressure according to the
high pressure difference of the non-azeotropic mixed
refrigerant.
[0025] The non-azeotropic mixed refrigerant may include at least
two hydrocarbons. The at least two hydrocarbons may include at
least one first hydrocarbon selected from an upper group having an
evaporation temperature of -12.degree. C. or more at 1 bar and at
least one second hydrocarbon selected from a middle group having an
evaporation temperature of -50.degree. C. or more and less than
-12.degree. C. at 1 bar. A gliding temperature difference may be
4.degree. C. or more. It is possible to increase cycle efficiency
of the refrigerating system in which the non-azeotropic mixed
refrigerant is used and improve operation stability of the
refrigerating system.
[0026] A weight ratio of the at least one first hydrocarbon may be
50% or more. Therefore, compression work of the compressor may be
optimized.
[0027] The linear compressor may include a piston configured to
reciprocate to compress the non-azeotropic mixed refrigerant, and a
cylinder configured to guide the piston. Oil pumped by a pressure
difference of the non-azeotropic mixed refrigerant may be present
on a contact surface between the piston and the cylinder.
Therefore, abnormal supply of oil may be prevented, and lubrication
of the piston may be performed more smoothly.
[0028] According to embodiments disclosed herein, a refrigerating
apparatus may include a compressor configured to compress a
non-azeotropic mixed refrigerant, a condenser configured to
condense the refrigerant compressed by the compressor, an expander
configured to expand the refrigerant condensed by the condenser,
and an evaporator configured to evaporate the refrigerant expanded
by the expander and provide cold air to an inner space of the
refrigerating apparatus. The compressor may be selectively operated
in an intermittent operation mode and a continuous operation mode,
and in the continuous operation mode, the compressor may be
controlled to operate continuously even when a temperature of an
inner space of the refrigerating apparatus is within a target
temperature range. In the case of the non-azeotropic mixed
refrigerant, a large pressure difference may be obtained.
Therefore, components operated during operation of the compressor
in the continuous operation mode may operate more reliably without
external influence due to frictional force.
Advantageous Effects
[0029] According to embodiments disclosed herein, it is possible to
obtain a refrigerating apparatus capable of obtaining high
efficiency when a non-azeotropic mixed refrigerant is used.
DESCRIPTION OF DRAWINGS
[0030] FIG. 1 is a schematic temperature graph of a non-azeotropic
mixed refrigerant and air in a counterflow evaporator;
[0031] FIG. 2 is a graph showing a temperature difference between
an inlet and an outlet of an evaporator and a gliding temperature
difference of a non-azeotropic mixed refrigerant according to
compositions of isobutane and propane;
[0032] FIG. 3A is a graph showing a refrigeration cycle when
isobutane is used as a refrigerant; and
[0033] FIG. 3B is a graph showing a refrigeration cycle when a
non-azeotropic mixed refrigerant is used as a refrigerant;
[0034] FIG. 4 is a schematic view of a refrigerating apparatus
according to an embodiment;
[0035] FIG. 5 a cross-sectional view of a linear compressor applied
to a refrigerating apparatus, when a piston is retracted, according
to an embodiment;
[0036] and
[0037] FIG. 6 is a cross-sectional view of a linear compressor
applied to a refrigerating apparatus, when a piston is advanced,
according to an embodiment;
[0038] FIG. 7 is a cross-sectional view of an oilless linear
compressor according to an embodiment;
[0039] FIG. 8 is a cross-sectional view of a suction muffler
according to an embodiment;
[0040] FIG. 9 is a view showing a state in which a first filter is
coupled to a suction muffler, according to an embodiment;
[0041] FIG. 10 is a view showing a configuration around a
compression chamber according to an embodiment;
[0042] FIG. 11 is an exploded perspective view showing a state in
which a cylinder and a frame are coupled to each other, according
to an embodiment;
[0043] FIG. 12 is an exploded perspective view of a cylinder and a
frame according to an embodiment;
[0044] FIG. 13 is an exploded perspective view of a frame according
to an embodiment;
[0045] FIG. 14 is a cross-sectional view showing a state in which a
cylinder and a frame are coupled to each other according to an
embodiment;
[0046] FIG. 15 is a view of a cylinder according to an
embodiment;
[0047] FIG. 16 is an enlarged cross-sectional view of portion "A"
in FIG. 14;
[0048] FIG. 17 is a cross-sectional view showing a state in which a
frame and a cylinder are coupled to each other according to an
embodiment;
[0049] FIG. 18 is an enlarged view of portion "B" of FIG. 17;
[0050] FIG. 19 is a cross-sectional view showing refrigerant flow
of a linear compressor according to an embodiment;
[0051] FIG. 20 is a view showing flow of refrigerant discharged
from a compression chamber in first and second passages according
to an embodiment; and
[0052] FIG. 21 is a view showing a flow of a refrigerant in a third
passage.
BEST MODE
[0053] Hereinafter, embodiments will be described with reference to
the accompanying drawings. The embodiments are not limited to the
embodiments discussed hereinafter, and those skilled in the art who
understand the spirit will be able to easily propose other
embodiments falling within the scope by adding, modifying, and
deleting components. However, this also falls within the
spirit.
[0054] In the following description, contents are divided into
technical elements and described. First, a process of selecting a
type of a non-azeotropic mixed refrigerant will be described.
[0055] Selection of Type of Non-Azeotropic Mixed Refrigerant
[0056] Refrigerants to be mixed, which are suitable for the
non-azeotropic mixed refrigerant, are proposed. As the refrigerant
to be mixed, a hydrocarbon-based (HC-based) refrigerant may be
selected. Hydrocarbon-based refrigerant is an eco-friendly
refrigerant having a low ozone depletion potential (ODP) and a low
global warming potential (GWP). The criteria for selecting a
refrigerant suitable for the non-azeotropic mixed refrigerant among
the hydrocarbon-based refrigerants may be summarized as
follows.
[0057] First, from a viewpoint of compression work, when a
difference (pressure difference (.DELTA.P)) between a condensing
pressure (Pd or p1) and an evaporation pressure (Ps or p2) is
smaller, the compression work of the compressor is further reduced,
which is advantageous for efficiency. Therefore, refrigerants
having a low condensing pressure and a high evaporation pressure
may be selected. However, considering reliability of compressors,
an evaporation pressure of 50 kPa or more may be selected.
[0058] Second, from a viewpoint of utilization of production
facilities, refrigerants may be selected which have been used in
the past for compatibility of existing facilities and components.
Third, from a viewpoint of purchase costs of refrigerants,
refrigerants obtainable at low cost may be selected. Fourth, from a
viewpoint of safety, refrigerants that are not harmful to humans
when refrigerant leaks may be selected.
[0059] Fifth, from a viewpoint of reducing irreversible loss,
reduction of a temperature difference between a refrigerant and
cold air so as to increase efficiency of a cycle is desirable.
Sixth, from a viewpoint of handling, refrigerants that can be
conveniently handled at a time of work and may be conveniently
injected by handlers may be selected.
[0060] The above criteria for selecting refrigerants is variously
applied in selecting the non-azeotropic mixed refrigerant.
[0061] Classification and Selection of Hydrocarbons
[0062] Based on evaporation temperature (Tv), candidate
refrigerants suggested by the National Institute of Standards and
Technology are classified into three (upper, middle, and lower)
groups in descending order of evaporation temperature. A density of
refrigerant is higher as evaporation temperature increases.
[0063] A combination of candidate refrigerants capable of
exhibiting an evaporation temperature of -20.degree. C. to
-30.degree. C. suitable for the environment of refrigerating
apparatuses may be selected. Hereinafter, classification of the
candidate refrigerants will be described.
[0064] The candidate refrigerants are classified into three types
based on boundary values of evaporation temperature, that is,
-12.degree. C. and -50.degree. C. The candidate refrigerants
classified into the three types are shown in Table 1. It can be
seen that the classification of the evaporation temperature changes
greatly based on the boundary values.
TABLE-US-00001 TABLE 1 Evaporation Evaporation temperature
temperature Triple point Hydrocarbon (1 bar) (20 bar) temperature
No. group name .degree. C. 1 upper isopentane 27.5 154.7 -159.85 2
1,2- 10.3 124.8 -136.25 butadiene 3 n-butane -0.9 114.5 -138.25 4
butene -6.6 105.8 -185.35 5 isobutane -12 100.7 -159.65 6 middle
propadiene -34.7 68.2 -136.25 7 propane -42.4 57.3 -187.71 8
propylene -47.9 48.6 -185.26 9 lower ethane -88.8 -7.2 -182.80 10
ethylene -104 -29.1 -169.15
[0065] Referring to Table 1, refrigerants that may be mixed as the
non-azeotropic mixed refrigerant may be selected and combined in
each region. First, which group is selected among the three groups
will be described. There may be one case in which refrigerants are
selected from the three groups and three refrigerants are mixed,
and three cases in which refrigerants are selected from two groups
and two refrigerants are mixed.
[0066] When at least one refrigerant is selected from each of the
three groups and three or more refrigerants are mixed, the
temperature rise and drop in the non-azeotropic mixed refrigerant
may be excessively great. In this case, design of the refrigerating
system may be difficult.
[0067] Thus, the non-azeotropic mixed refrigerant may be obtained
by selecting at least one refrigerant from each of two groups. At
least one refrigerant may be selected from each of the middle group
and the lower group, from each of the upper group and the middle
group, and from each of the upper group and the lower group. Among
them, a composition in which at least one refrigerant selected from
each of the upper group and the middle group is mixed may be
provided as the non-azeotropic mixed refrigerant.
[0068] When at least one refrigerant selected from each of the
middle group and the lower group is mixed, the evaporation
temperature of the refrigerant is excessively low. Thus, a
difference between interior temperature and the evaporation
temperature of the refrigerant is excessively great in a general
refrigerating apparatus. Therefore, efficiency of the refrigeration
cycle deteriorates and power consumption increases.
[0069] When at least one refrigerant selected from each of the
upper group and the lower group is mixed, a difference in
evaporation temperature between the at least two refrigerants is
excessively great. Therefore, unless a special high-pressure
environment is created, each refrigerant is classified into a
liquid refrigerant and a gaseous refrigerant under actual use
conditions. For this reason, it is difficult to inject the at least
two refrigerants together into a refrigerant pipe.
[0070] Selection of Hydrocarbons in Groups of Hydrocarbons
[0071] Which refrigerant is selected from the upper group and the
middle group will be described hereinafter.
[0072] First, the refrigerant selected from the upper group will be
described. At least one refrigerant selected from the upper group
may be used as the non-azeotropic mixed refrigerant.
[0073] As isopentane and butadiene have a relatively high
evaporation temperature, the inner temperature of the evaporator of
the refrigerating apparatus is limited and freezing efficiency
deteriorates. Isobutane and N-butane may be used without changing
components of the refrigeration cycle, such as the compressor of
the refrigerating apparatus, currently used. Therefore, their use
is most expected among the refrigerants included in the upper
group.
[0074] N-butane has a smaller compression work than isobutane, but
has a low evaporation pressure (Ps), which may cause a problem in
the reliability of the compressor. For this reason, isobutane may
be selected from the upper group. As described above, selection of
at least one from the other hydrocarbons included in the upper
group is permissible.
[0075] The refrigerant selected from the middle group will be
described hereinafter. At least one refrigerant selected from the
middle group may be used in the non-azeotropic mixed
refrigerant.
[0076] As propadiene has a smaller pressure difference (.DELTA.P)
than that of propane, efficiency is high. However, propadiene is
expensive and harmful to respiratory systems and skin when humans
inhale due to leakage. Propylene has a greater pressure difference
than that of propane, and thus, compression work of the compressor
is increased.
[0077] For this reason, propane may be selected from the middle
group. As described above, selection of at least one from the other
hydrocarbons included in the middle group is permissible.
[0078] For reference, isobutane may also be referred to as R600a,
and propane may also be referred to as R290. Although isobutane and
propane may be selected, other hydrocarbons belonging to the same
group may be applied in obtaining properties of the non-azeotropic
mixed refrigerant, even where there is no specific mention in the
following description. For example, if it is possible to obtain a
similar gliding temperature difference of the non-azeotropic mixed
refrigerant, other compositions than isobutane and propane may be
used.
[0079] Selection of Ratio of Selected Hydrocarbon Refrigerant,
Considering Power Consumption of Compression Work
[0080] As the refrigerant to be mixed in the non-azeotropic mixed
refrigerant, isobutane is selected from the upper group and propane
is selected from the middle group. Ratios of the refrigerants to be
mixed in the non-azeotropic mixed refrigerant may be selected as
follows.
[0081] Power consumption of the compressor, which is a main energy
consumption source of the refrigerating system, depends on pressure
difference. In other words, as the pressure difference increases,
more compression work need to be consumed. As the compression work
increases, efficiency of the cycle further deteriorates.
[0082] Isobutane has a smaller pressure difference (.DELTA.P) than
that of propane. For this reason, the non-azeotropic mixed
refrigerant may be provided with a weight ratio of isobutane of 50%
or more and a weight ratio of propane of 50% or less.
[0083] In the case of a composition in which the non-azeotropic
mixed refrigerant includes isobutane and propane mixed at a ratio
of 5:5, the condensing pressure is 745.3 kPa, the evaporation
pressure is 120.5 kPa, and the pressure difference is 624.7 kPa. In
the case of a composition in which the non-azeotropic mixed
refrigerant is substantially isobutane with a very small amount of
propane, the condensing pressure is 393.4 kPa, the evaporation
pressure is 53.5 kPa, and the pressure difference is 340.0 Pa.
[0084] The pressure is obtained by measuring an average value when
the compressor is turned on under ISO power consumption measurement
conditions. All values related to the composition of the
non-azeotropic mixed refrigerant are obtained under the same
conditions.
[0085] Ranges of the condensing pressure, the evaporation pressure,
and the pressure difference of the non-azeotropic mixed refrigerant
may be known using a mixing ratio of isobutane to propane that can
reduce the compression work as described above.
[0086] Selection of Ratio of Selected Hydrocarbon Refrigerant,
Considering Irreversible Loss of Evaporator
[0087] As described above, the non-azeotropic mixed refrigerant has
a gliding temperature difference (GTD) upon phase change. Using the
gliding temperature difference, evaporators may be sequentially
installed in a freezer compartment and a refrigerating compartment
to provide an appropriate temperature atmosphere for each
partitioned space. According to the gliding temperature difference,
a temperature difference between air and refrigerant evaporated in
each evaporator may be reduced, thereby reducing irreversibility
occurring during heat exchange. Reduction in irreversible loss may
reduce the loss of the refrigerating system.
[0088] FIG. 1 is a schematic temperature graph of a non-azeotropic
mixed refrigerant and air in a counterflow evaporator. In FIG. 1,
the horizontal axis represents progress distance, and the air and
the non-azeotropic mixed refrigerant move in opposite directions as
indicated by arrows. In FIG. 1, the vertical axis represents
temperature. Referring to FIG. 1, 1 is a line for air, 2 is a line
for the non-azeotropic mixed refrigerant, 3 is a line for
temperature rise of the non-azeotropic mixed refrigerant, 4 is a
line for temperature drop of the non-azeotropic mixed refrigerant,
and 5 is a line for a single refrigerant. Referring to the line 1
for air, for example, the temperature of the air may drop from a
range of -20.degree. C. to -18.degree. C. and the air may pass
through the evaporator.
[0089] Referring to the line 2 for the non-azeotropic mixed
refrigerant, the temperature of the non-azeotropic mixed
refrigerant may rise from -27.degree. C. and the non-azeotropic
mixed refrigerant may pass through the evaporator. The gliding
temperature difference of the non-azeotropic mixed refrigerant may
change according to the ratio of isobutane to propane. When the
gliding temperature difference is increased, the line 2 for the
non-azeotropic mixed refrigerant may move toward the line 3 for the
temperature rise of the non-azeotropic mixed refrigerant. When the
gliding temperature difference is decreased, the line 2 for the
non-azeotropic mixed refrigerant may move toward the line 4 for the
temperature drop of the non-azeotropic mixed refrigerant. For
reference, as there is no phase change in the single refrigerant,
there is no temperature change in the line 5 for the single
refrigerant.
[0090] Irreversible loss when heat exchange occurs cannot be
avoided due to the temperature difference between two interfaces
where heat exchange occurs. For example, when there is no
temperature difference between interfaces of two objects that
exchange heat with each other, there is no irreversible loss, but
heat exchange does not occur.
[0091] However, there are various methods for reducing irreversible
loss due to heat exchange. A representative method is to configure
a heat exchanger with counterflow. A counterflow heat exchanger may
reduce irreversible loss by allowing the temperature difference
between moving fluids to be reduced as much as possible.
[0092] In the case of an evaporator to which the non-azeotropic
mixed refrigerant is applied, the heat exchanger may be configured
with counterflow as shown in FIG. 1. As the temperature of the
non-azeotropic mixed refrigerant is increased during evaporation
due to the gliding temperature difference, the temperature
difference between the air and the non-azeotropic mixed refrigerant
may be reduced. When the gliding temperature difference of the
non-azeotropic mixed refrigerant and the temperature difference of
the air are reduced, irreversible loss may be reduced and
efficiency of the refrigeration cycle may be increased.
[0093] The gliding temperature difference of the non-azeotropic
mixed refrigerant may not be increased infinitely due to
limitations of the refrigerant. In addition, when the gliding
temperature difference of the non-azeotropic mixed refrigerant is
changed, the gliding temperature difference of the cold air is
changed. Accordingly, a size of the evaporator is changed and total
efficiency of the refrigeration cycle is affected. For example,
when the gliding temperature difference is increased, the inlet
temperature of the refrigerant is decreased or the outlet
temperature of the refrigerant is overheated, thus reducing
efficiency of the refrigeration cycle.
[0094] On the other hand, the gliding temperature difference of the
non-azeotropic mixed refrigerant and the temperature difference of
the air may converge to zero if a size of the heat exchanger is
infinitely large. However, considering mass productivity and cost
reduction of the heat exchanger, in the case of a general
refrigerating apparatus, the gliding temperature difference of the
non-azeotropic mixed refrigerant and the temperature difference of
the air are about 3.degree. C. to 4.degree. C.
[0095] FIG. 2 is a graph showing a temperature difference between
an inlet and an outlet of an evaporator and a gliding temperature
difference of a non-azeotropic mixed refrigerant according to
compositions of isobutane and propane. The horizontal axis
represents a content of isobutane, and the vertical axis represents
a temperature difference.
[0096] Referring to FIG. 2, when isobutane and propane are each
included in 100%, there is no temperature change while isobutane
and propane undergo evaporation as a single refrigerant. When
isobutane and propane are mixed, there are the gliding temperature
difference of the non-azeotropic mixed refrigerant and the
temperature difference between the inlet and the outlet of the
evaporator. A temperature difference 11 between the inlet and the
outlet of the evaporator is smaller than a gliding temperature
difference 12 of the non-azeotropic mixed refrigerant. This may be
caused by incomplete heat transfer between the refrigerant and
air.
[0097] When the gliding temperature difference of the
non-azeotropic mixed refrigerant is greater than the temperature
difference between the inlet and the outlet of the evaporator,
characteristics of the non-azeotropic mixed refrigerant may be well
utilized. Also, it is advantageous from a viewpoint of reducing
irreversibility in heat exchange and increasing efficiency of the
refrigeration cycle. Likewise, the gliding temperature difference
of the non-azeotropic mixed refrigerant may be greater than the
temperature difference of the air passing through the
evaporator.
[0098] In a general refrigerating apparatus, the temperature
difference of the air passing through the inlet and the outlet of
the evaporator may reach 4.degree. C. to 10.degree. C. In most
cases, the temperature difference of air is close to 4.degree. C.
For this reason, the gliding temperature difference of the
non-azeotropic mixed refrigerant may be maintained higher than
4.degree. C. Maintaining the gliding temperature difference to be
at least 4.1.degree. C. or higher, which is minimally higher than
the temperature difference between the inlet and the outlet of the
evaporator, may be advantageous. When the gliding temperature
difference of the non-azeotropic mixed refrigerant is less than
4.1.degree. C., thermal efficiency of the refrigeration cycle may
decrease.
[0099] In contrast, when the gliding temperature difference of the
non-azeotropic mixed refrigerant is greater than 4.1.degree. C.,
the temperature difference between the air and the refrigerant at
the outlet side of the refrigerant decreases, irreversibility
decreases, and thermal efficiency of the refrigeration cycle
increases. That the temperature difference between the air and the
refrigerant at the outlet side of the refrigerant decreases means
that the line 2 for the non-azeotropic mixed refrigerant moves
toward the line 3 for the temperature rise of the non-azeotropic
mixed refrigerant in FIG. 1.
[0100] In FIG. 2, when the gliding temperature difference of the
non-azeotropic mixed refrigerant is 4.1.degree. C., isobutane is
90%, and when the gliding temperature difference of the
non-azeotropic mixed refrigerant is greater than 4.1.degree. C.,
isobutane is less than 90%. In order to minimize compression work
of the compressor, isobutane may be 50% or more.
[0101] As a result, a weight ratio of the non-azeotropic mixed
refrigerant provided as isobutane and propane may be expressed as
in Equation 1.
50%.ltoreq.isobutane.ltoreq.90% [Math FIG. 1]
[0102] Propane is the remaining or other component in the weight
ratio of the non-azeotropic mixed refrigerant.
[0103] As the gliding temperature difference of the non-azeotropic
mixed refrigerant increases, irreversible loss may be reduced.
However, when the gliding temperature difference is excessively
great, a size of the evaporator becomes excessively large in order
to secure a sufficient heat exchange passage between the
refrigerant and the air. A space inside of the refrigerating
apparatus may be secured when the evaporator applied to a general
household refrigerating apparatus is designed with a capacity of
200 W or less. For this reason, the gliding temperature difference
of the non-azeotropic mixed refrigerant may be limited to
7.2.degree. C. or less.
[0104] In addition, when the gliding temperature difference of the
non-azeotropic mixed refrigerant is excessively great, the
temperature of the inlet of the evaporator may be too low or the
outlet of the evaporator outlet may be overheated too quickly,
based on the non-azeotropic mixed refrigerant. An available area of
the evaporator may be reduced and efficiency of the heat exchange
may decrease.
[0105] At the outlet of the evaporator, the temperature of the
non-azeotropic mixed refrigerant has to be higher than the
temperature of the air introduced into the evaporator. Otherwise,
efficiency of the heat exchanger decreases due to reversal of the
temperatures of the refrigerant and air. When this condition is not
satisfied, efficiency of the refrigerating system may be
reduced.
[0106] In FIG. 2, when the gliding temperature difference of the
non-azeotropic mixed refrigerant is 7.2.degree. C., isobutane is
75%, and when the gliding temperature difference of the
non-azeotropic mixed refrigerant is less than 7.2.degree. C.,
isobutane is more than 75%. As a result, considering this condition
and the condition of Equation 1 together, a weight ratio of the
non-azeotropic mixed refrigerant provided as isobutane and propane
may be expressed as in Equation 2.
75%.ltoreq.isobutane.ltoreq.90% [Math FIG. 2]
[0107] Propane is the remaining or other component in the weight
ratio of the non-azeotropic mixed refrigerant.
[0108] Selection of Ratio of Selected Hydrocarbon Refrigerant,
Considering Compatibility of Production Facilities and
Components
[0109] The temperature difference between the inlet and the outlet
of the evaporator of a general refrigerating apparatus may be set
to 3.degree. C. to 5.degree. C. This is due to various factors,
such components of the refrigerating apparatus, internal volume of
the machine room, heat capacity of each component, and size of the
fan, for example. When a composition ratio of the non-azeotropic
mixed refrigerant capable of providing the temperature of the inlet
and the outlet of the evaporator, that is, 3.degree. C. to
5.degree. C., is found in FIG. 2, it can be seen that isobutane is
between 76% and 87%.
[0110] As a result of the above discussion, the non-azeotropic
mixed refrigerant that satisfies all of the above-described
conditions may be expressed as Equation 3.
76%.ltoreq.isobutane'87% [Math FIG. 3]
[0111] Propane is the remaining or other component in the weight
ratio of the non-azeotropic mixed refrigerant.
[0112] Ratio of hydrocarbon refrigerant to be finally applied
[0113] The isobutane application range that can be selected on the
basis of the various criteria described above may be determined to
be 81% to 82%, which is the middle range of Equation 3. Propane may
occupy the remaining portion or component of the non-azeotropic
mixed refrigerant.
[0114] The case of using only isobutane was compared with the case
of using the non-azeotropic mixed refrigerant in which 85% of
isobutane and 15% of propane were applied. In both cases, the
evaporators were constructed in parallel to form the cycle of the
refrigerating system.
[0115] The experimental conditions were -29.degree. C. and
-15.degree. C. and the inlet temperatures of the compressors were
25.degree. C., respectively. Due to the difference in the
refrigerant, the temperature of the condenser was 31.degree. C.
when using only isobutane and 29.degree. C. when using the
non-azeotropic mixed refrigerant.
[0116] FIGS. 3A and 3B are tables for comparison of the
refrigeration cycle in each case. FIG. 3A is a graph showing the
refrigeration cycle when only isobutane is used. FIG. 3B is a graph
showing the refrigeration cycle when the non-azeotropic mixed
refrigerant is used.
[0117] In the experiment according to FIGS. 3A-3B, it can be seen
that when the non-azeotropic mixed refrigerant is used, improvement
in coefficient of performance was approximately 4.5%.
[0118] FIG. 4 is a schematic view of a refrigerating apparatus
according to an embodiment. Referring to FIG. 4, a refrigerating
apparatus according to an embodiment may include a machine room
631, a freezer compartment 632, and a refrigerating compartment
633. The refrigerating apparatus forms a refrigeration cycle that
operates the non-azeotropic mixed refrigerant. In the refrigeration
cycle, a compressor 621 that compresses the refrigerant, an
expander 622 that expands the compressed refrigerant, a condenser
623 that condenses the expanded refrigerant, and evaporators 624
and 625 may be included.
[0119] The compressor 621, the expander 622, and the condenser 623
may be provided in the machine room 631. The first evaporator 624
may be provided in the freezer compartment 632. The second
evaporator 625 may be provided in the refrigerating compartment
633. The freezer compartment and the refrigerating compartment may
be referred to as an "interior space".
[0120] The non-azeotropic mixed refrigerant is lower in the first
evaporator 624 than in the second evaporator 625. As the first
evaporator 624 is placed in the freezer compartment 632, the
refrigeration cycle may be operated more appropriately in the
partitioned space of the refrigerating apparatus. Therefore,
irreversible loss may be further reduced in an evaporation
operation of the evaporator 625.
[0121] Hereinafter, a linear compressor applicable as compressor
621 provided in the refrigerating apparatus will be described.
[0122] FIG. 5 a cross-sectional view of a linear compressor applied
to a refrigerating apparatus, when a piston is retracted, according
to an embodiment. FIG. 6 is a cross-sectional view of a linear
compressor applied to a refrigerating apparatus, when a piston is
advanced, according to an embodiment.
[0123] In the linear compressor according to this embodiment, as
shown in FIGS. 5 and 6, a linear compressor 60 is installed so as
to be buffered inside a shell 50. The shell 50 may include a lower
shell 51 with an open upper portion and an upper shell 52 mounted
to cover an upper side of the lower shell 51. An airtight space may
be defined between the lower shell 51 and the upper shell 52, and
oil O may be contained in an inner lower portion of the lower shell
51.
[0124] The oil allows a piston to be lubricated at an interface of
a cylinder. When lubrication is not smooth, it may cause a big
problem in reliability of the linear compressor.
[0125] The shell 50 may include a suction pipe 53 through which
fluid may be suctioned that passes therethrough, and a discharge
pipe 54 through which the fluid compressed by linear compressor 60
may be discharged that passes therethrough. The linear compressor
60 may include a damper 55 installed in the lower shell 51 to
provide vibrational support. The linear compressor 60 may include a
cylinder block 66 provided with a cylinder 62, a back cover 74
provided with a fluid suction pipe 72, a piston 80 provided to
linearly reciprocate in the cylinder 62 and including a fluid
suction passage 78 and a suction port 79 formed such that the fluid
is suctioned into the cylinder 62, a suction valve 82 installed in
the piston 80 to open and close the fluid suction passage 78, a
linear motor 84 that linearly reciprocates the piston 80, and a
discharge valve assembly 92 that forms a compression chamber C
between the discharge valve assembly 92 and the piston 80. When the
fluid inside the compression chamber C is compressed to a
predetermined pressure or higher, the compressed fluid is
discharged to the discharge pipe 54, which may be a loop pipe.
[0126] The cylinder 62 may be installed at a center of the cylinder
block 66. The back cover 74 may be mounted on a stator cover 152,
which is described hereinafter, as a coupling member, such as a
coupling bolt.
[0127] A flange 81 may be formed at a rear end of the piston 80.
The flange 81 may be connected to the linear motor 84 by a coupling
member, such as a coupling bolt, to receive drive force of the
linear motor 84.
[0128] The suction valve 82 may be an elastic member coupled to a
front end surface of the piston 80 by the coupling bolt and open
and close the suction port 79 by a pressure difference between the
compression chamber C and the suction port 79.
[0129] The linear motor 84 may include an outer core 85 installed
in the cylinder block 64, a bobbin 86 installed in the outer core
85, a coil 87 wound around the bobbin 86, an inner core 88
installed in the cylinder block 64 so as to have a predetermined
gap with the outer core 85, a magnet 89 located between the outer
core 85 and the inner core 88 so as to reciprocate linearly by the
electromagnetic force formed by the coil 87, and a magnet frame 90
in which the magnet 89 is mounted and which is coupled to the
flange 81 of the piston 80 to transmit a linear motion to the
piston 80.
[0130] The discharge valve assembly 92 may include a discharge
valve 93 that opens and closes a front end of the cylinder 62, an
inner discharge cover 96 in which the discharge valve 93 may be
elastically supported by discharge spring 94 and a fluid discharge
hole 95 may be formed, an outer discharge cover 97 in which a
passage may be formed between the outer discharge cover 97 and the
inner discharge cover 96, and a connection pipe 98 mounted on the
outer discharge cover 97 and connected to the discharge pipe
54.
[0131] The linear compressor 60 may be provided with an oil suction
passage 10 such that the oil O contained in the shell 50 may flow
between the cylinder 62 and the piston 80. An oil discharge passage
20 may be provided such that the oil O between the cylinder 62 and
the piston 80 may be discharged. As shown in FIG. 5 when the piston
80 is retracted, a low pressure is formed between the cylinder 62
and the piston 80. As shown in FIG. 6, a pump 30 that forms a high
pressure is provided between the cylinder 62 and the piston 80.
[0132] As described above, the pump 30 may be operated by a
pressure difference between a low pressure and a high pressure
formed between the cylinder 62 and the piston 80. When the
non-azeotropic mixed refrigerant is used, the pressure difference
between the low pressure and the high pressure formed between the
cylinder and the piston is large. Therefore, more oil may be pumped
with a greater force, as compared to a refrigerating apparatus
using a single refrigerant provided with equal cooling power.
[0133] The description of the non-azeotropic mixed refrigerant and
the pump 30 will be described hereinafter with operation of the
linear compressor.
[0134] The oil suction passage 10 may include an oil pipe 11
immersed in the oil O contained in the shell 50 and mounted on the
cylinder block 66, an oil cover 13 that communicates with the oil
pipe 11 and having an oil passage 12 formed between the oil cover
13 and the cylinder block 66, a cylinder block suction passage 14
formed in the cylinder block 66 such that oil suctioned into the
oil passage 12 passes through the cylinder block 66, and a cylinder
suction passage 15 formed in the cylinder 62 such that oil
suctioned into the cylinder block suction passage 14 is suctioned
into the inside of the pump 30. The oil discharge passage 20 may
include a cylinder discharge passage 21 formed in the cylinder 62
such that the oil O inside of the pump 30 may be discharged, and a
cylinder block discharge passage 22 formed in the cylinder block 66
such that the oil O discharged to the cylinder discharge passage 21
may be discharged through the cylinder block 66.
[0135] The pump 30 may include a cylinder stepped portion or step
31 formed on an inner circumferential surface of the cylinder 62,
and a piston stepped portion or step 32 formed on an outer
circumferential surface of the piston 80 to form a low pressure
between the piston stepped portion 32 and the cylinder stepped
portion 31 when the piston 80 is retracted, and form a high
pressure between the piston stepped portion 32 and the cylinder
stepped portion 31 when the piston 80 is advanced. An inclined
portion of the cylinder stepped portion 31 faces an inclined
portion of the piston stepped portion 32. An inner diameter D1 in
front of the inclined portion of the cylinder 62 is smaller than an
outer diameter D2 behind the inclined portion of the piston 80. The
cylinder stepped portion 31 and the piston stepped portion 32 form
a cylindrical space when the piston 80 is retracted.
[0136] The linear compressor 60 may further include an oil suction
valve 40 that opens one or a first side of the oil suction passage
10 by the low pressure formed in the pump 30 when the piston 80 is
retracted, and seals the one side of the oil suction passage 10
when the low pressure is released. The oil suction valve 40 may be
made of an elastic member fixed to the cylinder block 66 by a
coupling member, such as a coupling bolt, and a portion thereof may
be bent to open an inlet of the oil suction passage 10, in
particular, the oil passage 12.
[0137] The linear compressor 60 may further include an oil
discharge valve 150 that opens one or a first side of the oil
discharge passage 20 by the high pressure formed in the pump 30
when the piston 80 is retracted, and seals the one side of the oil
discharge passage 20 when the high pressure is released. The oil
discharge valve 150 may be made of an elastic member fixed to the
cylinder block 66 by a coupling member, such as a coupling bolt,
and a portion thereof may be bent to open an outlet of the oil
discharge passage 20, in particular, the cylinder block discharge
passage 22.
[0138] Reference numeral 152 represents a stator cover coupled to
the outer core 86 by a coupling member, such as a coupling bolt, to
cover a rear surface of the outer core 85. Reference numeral 27
represents a spring supporter, in which a first spring 29 may be
disposed between the spring supporter 27 and the back cover 72, and
a second spring 28 may be disposed between the spring supporter 27
and the stator cover 152. The spring supporter 27 may be fixed to
the flange 81 of the piston 80 by a coupling member, such as a
coupling bolt. Reference numeral 160 represents a muffler installed
on a rear end side of the piston 80 to guide the fluid suctioned
into the suction pipe 71 of the back cover 72 to the fluid suction
passage 78 of the piston 80 and reduce noise.
[0139] Operation of the linear compressor configured as described
above will be described hereinafter.
[0140] First, when a voltage is applied to the coil 87, a magnetic
field is formed around the coil 87. The magnet 89 linearly
reciprocates due to interaction with the magnetic field. The linear
reciprocating motion of the magnet 89 may be transmitted to the
piston 80 through the magnet frame 90, and the piston 80 may
reciprocate linearly in the cylinder 62.
[0141] The suction valve 82 and the discharge valve 93 may be
opened or closed by a pressure difference before and after
compression according to the linear reciprocating motion of the
piston 80. The fluid inside of the shell 50 may be suctioned into
the compression chamber C by sequentially passing through the fluid
suction pipe 72 of the back cover 74, the muffler 160, the fluid
suction passage 78, and the suction port 79 of the piston 80. The
fluid may be compressed by the piston 80 and discharged by
sequentially passing through the discharge valve assembly 92 and
the discharge pipe 54.
[0142] As described above, the piston 80 linearly reciprocates, and
while the fluid inside of the shell 50 is
suctioned/compressed/discharged, the oil O contained in the inner
lower portion of the shell 50 may be suctioned into the pump 30
according to a pressure change inside of the pump 30,
lubricate/cool between the cylinder 62 and the piston 80, and be
discharged outside of the linear compressor 60.
[0143] The pressure change of the pump 30 and the oil supply
process according to the pressure change will be described
hereinafter.
[0144] When the piston 80 is retracted as shown in FIG. 5, the
piston stepped portion 32 is positioned away from the cylinder
stepped portion 31. A low pressure is formed between the piston
stepped portion 32 and the cylinder stepped portion 31. The oil
suction valve 40 is partially bent by the low pressure to open the
inlet of the oil suction passage 10, in particular, the oil passage
12. The oil discharge valve 150 may seal the oil discharge passage
20, in particular, the cylinder block discharge passage 150 by the
low pressure. The oil O contained in the inner lower portion of the
shell 50 may pass through the oil pipe 11, the oil passage 12, the
cylinder block suction passage 14, and the cylinder suction passage
15 due to the low pressure, be suctioned into the space between the
piston stepped portion 32 and the cylinder stepped portion 31, and
lubricate/cool the cylinder 62 and the piston 80.
[0145] On the other hand, when the piston 80 is advanced as shown
in FIG. 6, the piston stepped portion 32 may be positioned close to
the cylinder stepped portion 31. A high pressure may be formed
between the piston stepped portion 32 and the cylinder stepped
portion 31. The oil suction valve 40 may seal the inlet of the oil
suction passage 10, in particular, the oil passage 12, due to the
high pressure. The oil discharge valve 150 may be partially bent to
open the oil discharge passage 20, in particular, the cylinder
block discharge passage 150. The oil in the space between the
piston stepped portion 32 and the cylinder stepped portion 31 may
pass through the cylinder discharge passage 21 and the cylinder
block discharge passage 22 due to the high pressure and be
discharged outside of the linear compressor 60.
[0146] The pump 30 may supply a large amount of oil with a greater
force as the pressure difference of the linear compressor
increases. The pressure difference of the linear compressor may
correspond to a pressure difference (.DELTA.P) between an
evaporation pressure and a condensation pressure of the
refrigerant.
[0147] When the non-azeotropic mixed refrigerant is used as the
working fluid of the linear compressor, it can operate with a
greater pressure difference (.DELTA.P) than when using isobutene
alone, which is widely used as a single refrigerant. The pressure
difference of the non-azeotropic mixed refrigerant applied to
embodiments will be described hereinafter.
[0148] First, in the case of a composition in which isobutane and
propane are mixed at a ratio of 5:5, the condensing pressure is
745.3 kPa, the evaporation pressure is 20.5 kPa, and the pressure
difference is 624.7 kPa. In the case of a composition in which the
non-azeotropic mixed refrigerant is substantially isobutane and a
very small amount of propane is mixed in, the condensing pressure
is 393.4 kPa, the evaporation pressure is 53.5 kPa, and the
pressure difference is 340.0 Pa.
[0149] As a result, the pressure difference of the non-azeotropic
mixed refrigerant of this embodiment may have a range of 340.0 kPa
or more and 624.7 kPa or less. As such, when the non-azeotropic
mixed refrigerant is used as the working fluid, oil may be supplied
to the linear compressor with a greater pressure difference. In
this case, it is possible to prevent a phenomenon in which the oil
supply passage is blocked, oil circulation is slowed, or a contact
surface between the piston and the cylinder is overheated.
[0150] When the non-azeotropic mixed refrigerant is used in the
linear compressor, a greater advantage may be expected when the
refrigerating system is operated in a continuous operation mode.
First, the continuous operation mode will be described
hereinafter.
[0151] The refrigerating apparatus is driven differently in the
refrigeration cycle according to a temperature region divided by an
upper limit reference value and a lower limit reference value based
on a set or predetermined temperature of an interior of the
refrigerating apparatus. More specifically, the temperature region
is divided into three types as follows. The temperature region may
be divided into a first case in which the interior temperature is
within a satisfaction temperature range falling between the upper
and lower limit reference values, a second case in which the
interior temperature is in a dissatisfaction temperature range
exceeding the upper limit reference value, and a third case in
which the interior temperature is in a supercooling temperature
range in which the interior temperature is below the lower limit
reference value.
[0152] A controller of the refrigerating apparatus may control the
refrigerating apparatus to operate to supply cooling power when the
interior temperature reaches the dissatisfaction temperature range,
and control the refrigerating apparatus to stop supplying cooling
power when the interior temperature reaches the supercooling
temperature range. Such an operation mode may be referred to as an
"intermittent operation mode".
[0153] When the upper limit reference value and the lower limit
reference value are set narrowly such that an amount of temperature
change in the refrigerating apparatus is reduced, a freshness of an
article may be improved and a storage period of the article may be
improved. Further, a constant temperature function may be improved
and the refrigerating apparatus may operate more suitably.
[0154] In the intermittent operation mode, the number of times the
refrigeration cycle is driven and stopped increases to improve the
constant temperature function. Due to this, reliability of
components is reduced by frequent on/off of refrigeration cycle
components, and power consumption increases each time the
refrigerating apparatus is switched from off to on. Also, when it
is switched from off to on, excessive cooling power is initially
supplied, and the constant temperature function deteriorates.
[0155] In order to solve this problem, the continuous operation
mode may be applied to reduce the number of times the refrigeration
cycle is driven and stopped. In the intermittent operation mode,
when the interior temperature reaches the upper limit reference
value and the lower limit reference value, the controller of the
refrigerating apparatus performs a switching control such that the
refrigerating apparatus starts or stops supplying cooling power.
When the interior temperature is in the satisfaction temperature
range, the controller controls the refrigerating apparatus to stop
supplying cooling power.
[0156] On the other hand, in the continuous operation mode, the
cooling power supply of the refrigerating apparatus is not stopped
even when the interior temperature is within the satisfaction
temperature range. The controller may control the refrigerating
apparatus based on a current temperature measured by a temperature
sensor and a target temperature. For example, the controller may
control the cooling power of the refrigerating apparatus based on
the difference between the target temperature and the current
temperature detected by the temperature sensor. As another example,
the controller may control the cooling power of the refrigerating
apparatus based on the increase or decrease in the interior
temperature detected by the temperature sensor at regular time
intervals.
[0157] The controller may control the cooling power of the
refrigerating apparatus based on the cooling power of the
refrigerating apparatus previously determined. For example, the
controller may determine the cooling power to be output at a
current time point as a value changing proportional to a sum of a
first output and a second output of the refrigerating apparatus
previously determined. As such, as the refrigerating apparatus is
operated in the continuous operation mode, the interior temperature
of the refrigerating apparatus may be maintained near the target
temperature without deviating to the dissatisfaction temperature
region.
[0158] The continuous operation mode may be performed in a state in
which an article is in a frozen or refrigerated state, and a
special operation may be performed in the intermittent operation
mode according to a special situation. For example, when a start
condition of the special operation in the refrigerating apparatus
is satisfied, the controller of the refrigerating apparatus may be
switched to the intermittent operation mode to drive the
refrigerating apparatus, and when an end condition of the special
operation is satisfied, the controller may switch to the continuous
operation mode to control operation of the refrigerating apparatus.
The special operation may include a defrosting operation, a door
load response operation, or an initial power application operation,
for example.
[0159] The defrosting operation will be described as an example.
When a defrosting cycle inside of the refrigerating apparatus has
elapsed, and/or when the evaporator temperature in the
refrigerating apparatus reaches a preset or predetermined specific
value, the controller may end the continuous operation mode and
switch to the intermittent operation mode. Operation of the
refrigerating apparatus in the defrosting operation will be
described hereinafter.
[0160] First, in order to pre-cool the internal temperature of the
refrigerating apparatus that will rise during the defrosting
operation, the controller may perform a deep cooling operation of
controlling the refrigerating apparatus to supply a greater cooling
power than that in the previous continuous operation mode. When a
predetermined time has elapsed after the deep cooling operation or
the temperature reaches a set or predetermined temperature, the
controller may end the deep cooling operation.
[0161] When the deep cooling operation ends, the controller may
stop supplying the cooling power of the refrigerating apparatus,
and the defrosting heater may be turned on to melt ice formed on
the evaporator. In this manner, the defrosting operation of
supplying heat to the refrigerating apparatus may be performed.
[0162] When the defrosting operation is completed, the interior
temperature of the refrigerating apparatus increased during the
defrosting operation has to be rapidly lowered. The controller may
allow the evaporator to receive a greater cooling power than that
received in the previous continuous operation mode. This may be
referred to as a "post-defrosting operation".
[0163] A point at which the post-defrosting operation ends is the
end condition of the defrosting operation. When the defrosting
operation ends or after the defrosting operation ends, the
controller may control the refrigerating apparatus to end the
intermittent operation mode and switch to the continuous operation
mode.
[0164] The door load response will be described as another
example.
[0165] When a predetermined time has elapsed after a door of the
refrigerating apparatus is opened or closed and/or when the
interior temperature of the refrigerating apparatus has reached a
preset or predetermined specific value after the door of the
refrigerating apparatus is opened or closed, a start condition of
the door load response operation is satisfied. At this time, the
controller controls the refrigerating apparatus to end the
continuous operation mode and switch to the intermittent operation
mode.
[0166] Operation of the refrigerating apparatus in the door load
response operation will be described hereinafter.
[0167] First, in order to remove a heat load introduced into the
interior of the refrigerating apparatus by the opening of the door,
the controller may perform the door load response operation of
controlling the refrigerating apparatus to supply a greater cooling
power than that in the previous continuous operation mode. When a
predetermined time has elapsed after the door load response
operation, or when the interior temperature of the refrigerating
apparatus reaches the preset or predetermined temperature, the
controller ends the door load response operation. When the door
load response operation ends or after the door load response
operation ends, the controller may control the refrigerating
apparatus to end the intermittent operation mode and switch to the
continuous operation mode.
[0168] The initial power application operation will be described as
another example.
[0169] When power is applied to the refrigerating apparatus again
after the power of the refrigerating apparatus is cut off, a start
condition of the initial power application operation is satisfied.
At this time, the controller may control the refrigerating
apparatus to operate in the intermittent operation mode rather than
the continuous operation mode.
[0170] More specifically, in order to rapidly lower the interior
temperature of the refrigerating apparatus increased while the
power is cut off, the controller performs an initial power
application operation of controlling the evaporator to supply a
greater cooling power than that in the previous continuous
operation mode. When a predetermined time has elapsed after the
initial power application operation, or when the interior
temperature of the refrigerating apparatus reaches the preset or
predetermined temperature, the controller ends the initial power
application operation.
[0171] When the initial power application operation ends or after
the initial power application operation ends, the controller may
control the refrigerating apparatus to end the intermittent
operation mode and switch to the continuous operation mode. When
operating in the continuous operation mode, the controller of the
refrigerating apparatus may control the evaporator to receive lower
cooling power than that received in the intermittent operation
mode. In particular, in the continuous operation mode that is
performed for a longer time than the intermittent operation mode,
the lower cooling power has to be supplied while being changed.
[0172] Such a refrigerating system that variably controls low
cooling power may be advantageous when the evaporation temperature
of the refrigerant is low. The relationship between the
refrigerating system and the evaporation temperature of the
refrigerant will be described.
[0173] Controlling the controller of the refrigerating apparatus
such that the cooling power of the evaporator is lowered means
reducing output of the compressor to reduce a flow rate or flow
amount of refrigerant circulating inside of a refrigerant pipe. In
a situation in which the refrigerant circulates continuously with a
low cooling power in order to perform the continuous operation
mode, in the refrigerating system in which the evaporation
temperature of the refrigerant is relatively high, a circulation
speed of the refrigerant in the refrigerating system in which the
evaporation temperature of the refrigerant is relatively low has to
be high in flow rate or flow amount of the refrigerant. This is
because when a same amount of refrigerant evaporates at a lower
temperature, more cooling power may be generated.
[0174] Consequently, compressor output in the refrigerating system
having a high evaporation temperature has to be greater than
compressor output in the refrigerating system having a low
evaporation temperature. Similarly, there is a disadvantage in that
power consumption for driving the compressor is increased. In
addition, as a period in which the continuous operation mode is
performed is longer, wear of components, such as the compressor,
may occur.
[0175] Due to the above advantages, it is advantageous that the
refrigerating apparatus to which the continuous operation mode is
applied uses the non-azeotropic mixed refrigerant mixed, in which a
refrigerant, such as propane, having a relatively low evaporation
temperature is mixed, rather than isobutane, which has a relatively
high evaporation temperature. In the refrigerating apparatus to
which the continuous operation mode is applied, reliability
problems may occur when the compressor is configured to cause the
piston of the compressor to be lifted in the cylinder by oil. As
seen in linear compression, linear compressors lubricated with oil
cause oil inside of the compressor to circulate due to the pressure
difference (.DELTA.P) of the refrigerant generated during a
reciprocating motion of the piston. Therefore, as at least one of a
frequency or stroke of the linear compressor is lowered during low
cooling power driving, the circulating speed of oil and the
circulating amount of oil decreases, and thus, an oil supply
passage may be blocked.
[0176] In the continuous operation mode, as there are many more
periods in which low cooling power is supplied than periods in
which high cooling power is supplied, problems may occur in the
circulation of the oil. In order to solve this problem, when the
non-azeotropic mixed refrigerant having a large pressure difference
(.DELTA.P) is used, it is possible to reduce the decrease in the
circulation speed of the oil and reduce the blocking of the oil
supply passage, as compared with a single refrigerant, such as
isobutane.
[0177] It has been described that the non-azeotropic mixed
refrigerant may be applied to the circulating operation of oil for
lubrication of the linear compressor. In addition, the
non-azeotropic mixed refrigerant may be applied even in the case of
a linear compressor in which oil is not used for lubrication. In
this case, the linear compressor may be lubricated by air. More
precisely, contact between the piston and the cylinder may be
prevented by air. Such a compressor may be referred to as an
"oilless linear compressor".
[0178] The oilless linear compressor may form a gas layer between
the cylinder and the piston instead of oil required to prevent wear
and damage between the cylinder and the piston. The oilless
compressor may lift the piston in the cylinder due to the pressure
difference (.DELTA.P) of the refrigerant generated during
reciprocation of the piston. Therefore, it is advantageous to apply
the non-azeotropic mixed refrigerant having a large pressure
difference (.DELTA.P).
[0179] FIG. 7 is a cross-sectional view of an oilless linear
compressor according to an embodiment. Referring to FIG. 7, the
linear compressor 100 according to this embodiment may include a
substantially cylindrical shell 101, a first cover 102 coupled to
one or a first side of the shell 101, and a second cover 103
coupled to the other or a second side of the shell 101. For
example, the linear compressor 100 may be laid in a horizontal
direction, the first cover 102 may be coupled to a first or right
side of the shell 101, and the second cover 103 may be coupled to a
second or left side of the shell 101. The linear compressor 100
exemplifies a linear compressor in which oil is not used. The first
cover 102 and the second cover 103 may be understood as components
of the shell 101.
[0180] The linear compressor 100 may include a cylinder 120
provided in the shell 101, a piston 130 that linearly reciprocates
in the cylinder 120, and a motor assembly 140 as a linear motor
that applies a drive force to the piston 130. When the motor
assembly 140 is driven, the piston 130 may reciprocate at a high
speed. An operating frequency of the linear compressor 100
according to this embodiment may be, for example, approximately 100
Hz.
[0181] The linear compressor 100 may include a suction inlet 104,
through which refrigerant may be introduced, and a discharge outlet
105, through which the refrigerant compressed in the cylinder 120
may be discharged. The suction outlet 104 may be coupled to the
first cover 102, and the discharge outlet 105 may be coupled to the
second cover 103.
[0182] The refrigerant suctioned through the suction outlet 104 may
flow through a suction muffler 150 to the piston 130. In a process
in which the refrigerant passes through the suction muffler 150,
noise may be reduced. The suction muffler 150 may be configured by
combining a first muffler 151 and a second muffler 153. At least a
portion of the suction muffler 150 may be located inside of the
piston 130.
[0183] The piston 130 may include a piston body 131 having an
approximately cylindrical shape, and a piston flange 132 that
extends from the piston body 131 in a radial direction. The piston
body 131 may reciprocate inside of the cylinder 120, and the piston
flange 132 may reciprocate outside of the cylinder 120.
[0184] The piston 130 may be made of a non-magnetic aluminum
material, for example, aluminum or an aluminum alloy. As the piston
130 is made of the aluminum material, a magnetic flux generated in
the motor assembly 140 may be transmitted to the piston 130 while
preventing leakage from the piston 130 to the outside. The piston
130 may be formed by a forging method, for example.
[0185] The cylinder 120 may be made of a non-magnetic aluminum
material, for example, aluminum or an aluminum alloy. A material
composition ratio of the cylinder 120 and the piston 130, that is,
a type and component ratio may be the same.
[0186] As the cylinder 120 is made of the aluminum material, the
magnetic flux generated in the motor assembly 200 may be
transmitted to the cylinder 120 while preventing leakage from the
cylinder 120 to the outside. The cylinder 120 may be formed by an
extrusion rod processing method, for example.
[0187] As the piston 130 and the cylinder 120 may be made of the
same material (aluminum), a coefficient of thermal expansion may be
the same. During operation of the linear compressor 100, an
environment of high temperature (about 100.degree. C.) may be
created inside of the shell 100. As the coefficient of thermal
expansion of the piston 130 and the cylinder 120 may be the same,
the piston 130 and the cylinder 120 may be thermally deformed by a
same amount. As a result, the piston 130 and the cylinder 120 may
be thermally deformed in different sizes or directions, thereby
preventing interference with the cylinder 120 during operation of
the piston 130.
[0188] The cylinder 120 may be configured to accommodate at least a
portion of the suction muffler 150 and at least a portion of the
piston 130. Inside the cylinder 120, a compression space P in which
the refrigerant is compressed by the piston 130 is formed. A
suction hole 133 through which the refrigerant may be introduced
into the compression space P may be defined on or at a front side
of the piston 130, and a suction valve 135 that selectively opens
the suction hole 133 may be disposed on or at a front side of the
suction hole 133. A coupling hole to which a predetermined coupling
member may be coupled may be defined in or at an approximately
central portion of the suction valve 135.
[0189] A discharge cover 160 that defines a discharge space 160a or
a discharge passage for the refrigerant discharged from the
compression space P, and a discharge valve assembly 161, 162, and
163 coupled to the discharge cover 160 to selectively discharge the
refrigerant compressed in the compression space P may be provided
at a front side of the compression space P. The discharge valve
assembly 161, 162, and 163 may include discharge valve 161 which
may be opened when the pressure of the compression space P is above
a discharge pressure to introduce the refrigerant into the
discharge space, valve spring 162 provided between the discharge
valve 161 and the discharge cover 160 to apply elastic force in an
axial direction, and stopper 163 that limits a deformation amount
of the valve spring 162. The compression space P may be understood
as a space defined between the suction valve 135 and the discharge
valve 161.
[0190] The "axial direction" may be understood as a direction in
which the piston 130 reciprocates, that is, the horizontal
direction in FIG. 7. Also, in the "axial direction", a direction
from the suction inlet 104 toward the discharge outlet 105, that
is, a direction in which the refrigerant flows may be defined as a
"frontward direction", and a direction opposite to the frontward
direction may be defined as a "rearward direction". On the other
hand, the "radial direction" may be understood as a direction
perpendicular to the direction in which the piston 130
reciprocates, that is, the vertical direction in FIG. 3.
[0191] The stopper 163 may be seated on the discharge cover 160,
and the valve spring 162 may be seated on or at a rear side of the
stopper 163. The discharge valve 161 may be coupled to the valve
spring 162, and a rear portion or a rear surface of the discharge
valve 161 may be supported on a front surface of the cylinder 120.
The valve spring 162 may include, for example, a plate spring. The
suction valve 135 may be disposed on or at one or a first side of
the compression space P, and the discharge valve 161 may be
disposed on the other or a second side of the compression space P,
that is, an opposite side of the suction valve 135.
[0192] While the piston 130 linearly reciprocates within the
cylinder 120, when the pressure of the compression space P is below
the discharge pressure and a suction pressure, the suction valve
135 may be opened to suction the refrigerant into the compression
space P. On the other hand, when the pressure of the compression
space P is above the suction pressure, the suction valve 135 may
compress the refrigerant of the compression space P in a state in
which the suction valve 135 is closed.
[0193] When the pressure of the compression space P is above the
discharge pressure, the valve spring 162 may be deformed to open
the discharge valve 161. The refrigerant may be discharged from the
compression space P into the discharge space 160a of the discharge
cover 160.
[0194] The refrigerant flowing through the discharge space of the
discharge cover 160 may be introduced into a loop pipe 165. The
loop pipe 165 may be coupled to the discharge cover 160 and extend
to the discharge outlet 105. The loop pipe 165 may guide the
compressed refrigerant of the discharge space to the discharge
outlet 105. For example, the loop pipe 165 may have a shape wound
in a predetermined direction, extending roundly, and coupled to the
discharge outlet 105.
[0195] The linear compressor 100 may further include a frame 110.
The frame 110 may fix the cylinder 120 and may be coupled to the
cylinder 120 by a separate coupling member. The frame 110 may
surround the cylinder 120. That is, the cylinder 120 may be
accommodated in the frame 110. The discharge cover 172 may be
coupled to a front surface of the frame 110.
[0196] At least a portion of the gas refrigerant among the high
pressure gas refrigerant discharged through the open discharge
valve 161 may flow toward an outer circumferential surface of the
cylinder 120 and a space between the cylinder 120 and the frame 110
which are coupled to each other. The refrigerant may be introduced
into the cylinder 120 through a gas inlet (122 in FIG. 16) and a
nozzle (123 in FIG. 16) formed in the cylinder 120. The introduced
refrigerant may flow into the space between the piston 130 and the
cylinder 120 such that an outer circumferential surface of the
piston 130 is spaced apart from an inner circumferential surface of
the cylinder 120. Therefore, the introduced refrigerant may
function as a "gas bearing" that reduces friction between the
cylinder 120 and the piston 130 during reciprocating motion.
[0197] The motor assembly 140 may include outer stators 141, 143,
and 145 fixed to the frame 110 and disposed to surround the
cylinder 120, an inner stator 148 spaced inward from the outer
stators 141, 143, and 145, and a permanent magnet 146 disposed in a
space between the outer stator 141, 143, and 145 and the inner
stator 148. The permanent magnet 146 may linearly reciprocate by
mutual electromagnetic force between the outer stators 141, 143,
and 145 and the inner stator 148. The permanent magnet 146 may be
provided as a single magnet having one polarity or may be provided
by coupling a plurality of magnets having three different
polarities.
[0198] The permanent magnet 146 may be coupled to the piston 130 by
a connecting member 138. The connecting member 138 may be coupled
to the piston flange 132 and bent and extended toward the permanent
magnet 146. When the permanent magnet 146 reciprocates, the piston
130 may reciprocate together with the permanent magnet 146 in the
axial direction.
[0199] The motor assembly 140 may further include a fixing member
147 that fixes the permanent magnet 146 to the connecting member
138. The fixing member 147 may be configured, for example, by
mixing a glass fiber or a carbon fiber with a resin. The fixing
member 147 may surround inner and outer sides of the permanent
magnet 146, such that a coupled state between the permanent magnet
146 and the connecting member 138 may be firmly maintained.
[0200] The outer stators 141, 143, and 145 may include coil winding
bodies 143 and 145, and stator core 141. The coil winding bodies
143 and 145 may include bobbin 143, and coil 145 wound in a
circumferential direction of the bobbin 143. A cross-section of the
coil 145 may have a polygonal shape, for example, a hexagonal
shape. The stator core 141 may be formed, for example, by
laminating a plurality of laminations in a circumferential
direction, and may surround the coil winding bodies 143 and
145.
[0201] A stator cover 149 may be disposed on one or a first side of
each of the outer stators 141, 143, and 145. The outer stators 141,
143, and 145 may each have one or a second side supported by the
frame 110 and the other side supported by the stator cover 149.
[0202] The inner stator 148 may be fixed to an outer circumference
of the frame 110. In the inner stator 148, a plurality of
laminations may be laminated in the circumferential direction
outside the cylinder 120.
[0203] The linear compressor 100 may further include a supporter
137 that supports the piston 130, and a back cover 170
spring-coupled to the supporter 137. The supporter 137 may be
coupled to the piston flange 132 and the connecting member 138 by a
predetermined coupling member.
[0204] At a front of the back cover 170, a suction guide 155 may be
provided. The suction guide 155 may guide refrigerant suctioned
through the suction inlet 104 into the suction muffler 150.
[0205] The linear compressor 100 may further include a plurality of
springs 176 adjusted in natural frequency to allow the piston 130
to perform a resonant motion. The plurality of springs 176 may
include a first spring supported between the support 137 and the
stator cover 149, and a second spring supported between the support
137 and the back cover 170.
[0206] The linear compressor 100 may further include plate springs
172 and 174 provided on or at both sides of the shell 101 such that
internal components of the compressor 100 may be supported by the
shell 101. The plate springs 172 and 174 include first plate spring
172 coupled to the first cover 102 and second plate spring 174
coupled to the second cover 103. For example, the first plate
spring 172 may be fitted to a portion at which the shell 101 and
the first cover 102 are coupled, and the second plate spring 174
may be disposed to fit in a portion at which the shell 101 and the
second cover 103 are coupled.
[0207] FIG. 8 is a cross-sectional view of a suction muffler
according to an embodiment. FIG. 9 is a view showing a state in
which a first filter is coupled to the suction muffler, according
to an embodiment.
[0208] Referring to FIGS. 8 and 9, the suction muffler 150 may
include first muffler 151, second muffler 153 coupled to the first
muffler 151, and a first filter 310 supported by the first muffler
151 and the second muffler 153. In the first muffler 151 and the
second muffler 153, a flow space through which the refrigerant may
flow is defined. The first muffler 151 may extend from an inside of
the suction inlet 104 toward the discharge outlet 105, and at least
a portion of the first muffler 151 may extend into the suction
guide 155. The second muffler 153 may extend from the first muffler
151 into the piston body 131.
[0209] The first filter 310 may be installed in the flow space to
filter foreign matter. The first filter 310 may be made of a
magnetic material, such that foreign matter included in the
refrigerant, in particular, metal dirt, may be easily filtered. For
example, the first filter 310 may be made of stainless steel and
may have a predetermined magnetic property, and a rust phenomenon
may occur. As another example, the first filter 310 may be coated
with a magnetic material or may be configured such that a magnetic
is attached to a surface of the first filter 310.
[0210] The first filter 310 may be a mesh type filter having a
plurality of filter holes and may have a substantially disc shape.
The filter holes may each have a diameter or a width of a
predetermined size or less. For example, the predetermined size may
be about 25 .mu.m.
[0211] The first muffler 151 and the second muffler 153 may be
assembled by a press-fitting method, for example. The first filter
310 may be assembled by being fitted into press-in portions of the
first muffler 151 and the second muffler 153.
[0212] A groove 151a to which at least a portion of the second
muffler 153 may be coupled may be formed in the first muffler 151.
The second muffler 153 may include a protrusion 153a inserted into
the groove 151a of the first muffler 151.
[0213] The first filter 310 may be supported by the first and
second mufflers 151 and 153 in a state in which both sides of the
first filter 310 are disposed between the groove 151a and the
protrusion 153a. In a state in which the first filter 310 is
located between the first and second mufflers 151 and 153, when the
first muffler 151 and the second muffler 153 move in a direction
closer to each other and are pressed, side or outer edge portions
of the first filter 310 may be fixed by being disposed between the
groove 151a and the protrusion 153a.
[0214] As such, by providing the first filter 310, foreign matter
having a predetermined size or more in the refrigerant suctioned
through the suction inlet 104 may be filtered by the first filter
310. Therefore, foreign matter is not included in the refrigerant
acting as the gas bearing between the piston 130 and the cylinder
120, thereby preventing the refrigerant from flowing into the
cylinder 120. In addition, as the first filter 310 is firmly fixed
to the press-in portions of the first and second mufflers 151 and
153, separation from the suction muffler 150 may be prevented.
[0215] In this embodiment, it has been described that the groove
151a is formed in the first muffler 151 and the protrusion 153a is
formed in the second muffler 153. However, the protrusion may be
formed in the first muffler 151 and the groove may be formed in the
second muffler 153.
[0216] FIG. 10 is a view showing a configuration around a
compression chamber according to an embodiment. FIG. 11 is an
exploded perspective view showing a state in which the cylinder and
the frame are coupled to each other, according to an embodiment.
FIG. 12 is an exploded perspective view showing a configuration of
the cylinder and the frame, according to an embodiment. FIG. 13 is
an exploded perspective view of the frame according to an
embodiment. FIG. 14 is a cross-sectional view showing a state in
which the cylinder and the frame are coupled to each other,
according to an embodiment.
[0217] Referring to FIGS. 10 to 14, in the linear compressor 100,
at least a portion of the refrigerant compressed and discharged in
the compression chamber P flows into a space between the frame 110
and the cylinder 120. The space between the frame 110 and the
cylinder 120 is understood as a gap formed between an inner surface
of the frame 110 and the outer surface of the cylinder 120 by an
assembly tolerance of the frame 110 and the cylinder 120. The space
between the frame 110 and the cylinder 120 may include passages
410, 420, and 430. The passages 410, 420 and 430 may include first
passage 410, second passage 420, and third passage 430, which may
be sequentially formed in a direction in which the refrigerant
flows.
[0218] The cylinder 120 may include an approximately cylindrical
cylinder body 121, and a cylinder flange 125 that extends radially
from the cylinder body 121. The cylinder body 121 may include gas
inlet 122 through which the discharged gas refrigerant may flow.
The gas inlet 122 may be formed in a circular shape along an outer
circumferential surface of the cylinder body 121.
[0219] A plurality of gas inlets 122 may be provided. The plurality
of gas inlets 122 may include gas inlets 122a and 122b (in FIG. 15)
located on one or a first side from an axial center of the cylinder
body 121, and gas inlet 122c (in FIG. 15) located on the other or a
second side from the axial center.
[0220] The cylinder flange 125 may be provided with a coupling
portion 126 coupled to the frame 110. The coupling portion 126 may
protrude from an outer circumferential surface of the cylinder
flange 125 in an outward direction. The coupling portion 126 may be
coupled to cylinder coupling hole 118 of the frame 110 by a
predetermined fastening member, for example, a bolt.
[0221] The cylinder flange 125 may include a seating surface 127
seated on the frame 110. The seating surface 127 may be a rear
portion of the cylinder flange 125 extending radially from the
cylinder body 121.
[0222] The frame 110 may include a frame body 111 that surrounds
the cylinder body 121, and a cover coupling portion 115 that
extends in a radial direction of the frame body 111 and coupled to
the discharge cover 160. The cover coupling portion 115 may be
provided with a plurality of cover coupling holes 116, into which
the coupling member coupled to the discharge cover 160 may be
inserted, and a plurality of cylinder coupling holes 118, into
which the coupling member coupled to the cylinder flange 125 may be
inserted The cylinder coupling hole 118 may be defined at a
somewhat recessed position from the cover coupling portion 115.
[0223] The frame 110 may be provided with a recessed portion or
recess 117. The recessed portion 117 may be recessed rearward from
the cover coupling portion 115, and the cylinder flange 125 may be
inserted into the recessed portion 117. That is, the recessed
portion 117 may surround an outer circumferential surface of the
cylinder flange 125. A recessed depth of the recessed portion 117
may correspond to a frontward-and-rearward width of the cylinder
flange 125.
[0224] A predetermined refrigerant flow space, that is, first
passage 410 may be formed between an inner circumferential surface
of the recessed portion 117 and the outer circumferential surface
of the cylinder flange 125. In a state in which the cylinder 120 is
assembled to the frame 110, a predetermined assembly tolerance may
be formed between the outer circumferential surface of the cylinder
flange portion 125 and the inner circumferential surface of the
recessed portion 117, and a space corresponding to the assembly
tolerance may define the first passage 410.
[0225] The high pressure gas refrigerant discharged from the
discharge valve 161 may flow through the first passage 410 into
second passage 420 provided with a second filter 320. The second
filter 320 may be a filter provided between the frame 110 and the
cylinder 120 to filter the high pressure gas refrigerant discharged
through the discharge valve 161.
[0226] A seating portion or seat 113 provided stepwise may be
formed at a rear end portion or end of the recessed portion 117.
The seating portion 113 may extend radially inward from the
recessed portion 117 and be located to face the seating surface 127
of the cylinder flange 125. A ring-shaped second filter 320 may be
seated on the seating portion 113.
[0227] In a state in which the second filter 320 is seated on the
seating portion 113, when the cylinder 120 is coupled to the frame
110, the cylinder flange 125 may press the second filter 320 in
front of the second filter 320. That is, the second filter 320 may
be disposed and fixed between the seating portion 113 of the frame
110 and the seating surface 127 of the cylinder flange 125.
[0228] The second passage 420 may be a passage through which the
refrigerant passing through the first passage 410 may flow. A
predetermined assembly tolerance may be formed between the seating
portion 113 and the seating surface 127 of the cylinder flange
portion 125, and the space corresponding to the assembly tolerance
may define the second passage 420.
[0229] The second filter 320 may be installed in the second passage
420 to prevent foreign matter in the high pressure gas refrigerant
flowing through the second passage 420 from flowing into the gas
inlet 122 of the cylinder 120, and adsorbs oil contained in the
refrigerant.
[0230] For example, the second filter 320 may include a nonwoven
fabric or an absorbent fabric made of polyethylene terephthalate
(PET) fibers. PET has excellent heat resistance and mechanical
strength. Further, PET can block foreign matter of 2 .mu.m or more
in the refrigerant.
[0231] Another embodiment is proposed hereinafter.
[0232] In the previous embodiment, the second filter 320 has been
described as being installed in the second passage 420. However,
the second filter 320 may be installed in the first passage 410,
that is, a space between the outer circumferential surface of the
cylinder flange 125 and the inner circumferential surface of the
recessed portion 117 of the frame 110.
[0233] The passages 410, 420 and 430 may include third passage 430
through which the refrigerant passing through the second passage
420 may flow. The third passage 430 may extend rearward from the
second passage 420 along the outer circumferential surface of the
cylinder body 121 and extend to a space between the rear portion of
the frame body 111 and the first body end (121a in FIG. 15) of the
cylinder body 121. The refrigerant flowing through the third
passage 430 may flow to the inner circumferential side of the
cylinder 120 via the gas inlet 122 and the nozzle 123.
[0234] FIG. 15 is a view showing of a cylinder according to an
embodiment. FIG. 16 is an enlarged cross-sectional view of portion
"A" in FIG. 14.
[0235] Referring to FIGS. 15 and 16, the cylinder 120 according to
this embodiment may include cylinder body 121 having a
substantially cylindrical shape and forming first body end 121a and
second body end 121b, and cylinder flange 125 that extends radially
outward from the second body end 121b of the cylinder body 121. The
first body end 121a and the second body end 121b form ends of the
cylinder body 121 based on axial center 121c of the cylinder body
121. The cylinder body 121 may include a plurality of gas inlets
122 through which at least a portion of the refrigerant of the high
pressure gas refrigerant discharged through the discharge valve 161
may flow, and a third filter 330 may be installed therein. The
cylinder body 121 may further include nozzle 123 that extends
radially inward from the plurality of gas inlets 122.
[0236] The plurality of gas inlets 122 and the nozzle 123 may be
one configuration of third passage 430. Therefore, at least a
portion of the refrigerant flowing through the third passage 430
may flow toward the inner circumferential surface of the cylinder
120 through the plurality of gas inlets 122 and the nozzle 123. The
plurality of gas inlets 122 may be recessed by a predetermined
depth and width from the outer circumferential surface of the
cylinder body 121.
[0237] The introduced refrigerant may flow between the outer
circumferential surface of the piston 130 and the inner
circumferential surface of the cylinder 120 and function as a gas
bearing. That is, due to the pressure of the refrigerant, the outer
circumferential surface of the piston 130 may maintain a state of
being spaced from the inner circumferential surface of the cylinder
120.
[0238] The plurality of gas inlets 122 may include first gas inlet
122a and second gas inlet 122b located at one or a first side from
the axial center 121c of the cylinder body 121, and a third gas
inlet 122c located at the other or a second side from the axial
center 121c. The first and second gas inlets 122a and 122b may be
located closer to the second body end 121b based on the axial
center 121c of the cylinder body 121, and the third gas inlet 122c
may be located closer to the first body end 121a based on the axial
center 121c of the cylinder body 121. That is, the plurality of gas
inlets 122 may be disposed in an asymmetrical number based on the
axial center 121c of the cylinder body 121.
[0239] Referring to FIG. 15, the internal pressure of the cylinder
120 may be higher at the second body end 121b closer to a discharge
side of the compressed refrigerant than the first body end 121a
closer to a suction side of the refrigerant. Therefore, more gas
inlets 122 may be formed toward the second body end 121b to enhance
the function of the gas bearing. However, relatively fewer gas
inlets 122 may be formed toward the first body end 121a.
[0240] The cylinder body 121 may further include nozzle 123 that
extends from the plurality of gas inlets 122 toward the inner
circumferential surface of the cylinder body 121. The nozzle 123
may have a smaller width or size than the gas inlet 122.
[0241] A plurality of nozzles 123 may be formed along the gas inlet
122 extending in a circular shape. The plurality of nozzles 123 may
be spaced apart from each other.
[0242] The nozzle 123 may include an inlet 123a connected to the
gas inlet 122 and an outlet 123b connected to the inner
circumferential surface of the cylinder body 121. The nozzle 123
may be may have a predetermined length from the inlet 123a toward
the outlet 123b.
[0243] A recessed depth and width of the plurality of gas inlets
122, and a length of the nozzle 123 may be determined to be an
appropriate size in consideration of a rigidity of the cylinder
120, an amount of the third filter 330, or a magnitude of a
pressure drop of the refrigerant passing through the nozzle 123.
For example, when the recessed depth and width of the plurality of
gas inlets 122 are too large or the length of the nozzle portion
123 is too small, the rigidity of the cylinder 120 may be weakened.
When the recessed depth and width of the plurality of gas inlet
portions 122 are too small, the amount of the third filter 330
installed in the gas inlet 122 may be too small. When the length of
the nozzle 123 is too large, the pressure drop of the refrigerant
passing through the nozzle 123 becomes too large. Thus, the
refrigerant cannot perform a sufficient function as a gas
bearing.
[0244] A diameter of the inlet 123a of the nozzle portion 123 may
be larger than a diameter of the outlet 123b. When the diameter of
the nozzle 123 is too large, the amount of the refrigerant flowing
into the nozzle 123 among the high pressure gas refrigerants
discharged through the discharge valve 161 becomes too large. Thus,
a flow rate loss of the compressor is increased. When the diameter
of the nozzle 123 is too small, the pressure drop in the nozzle 123
becomes large and performance of the gas bearing decreases.
Therefore, in this embodiment, the diameter of the inlet 123a of
the nozzle 123 is relatively large to reduce the pressure drop of
the refrigerant flowing into the nozzle 123, and the diameter of
the outlet 123b is formed relatively small such that the amount of
the gas bearing flowing through the nozzle 123 may be adjusted to a
predetermined value or less.
[0245] The third filter 330 may be installed in the plurality of
gas inlets 122. The third filter 330 may filter the refrigerant
flowing toward the inner circumferential surface of the cylinder
120.
[0246] The third filter 330 may prevent foreign matter having a
predetermined size or more from being introduced into the cylinder
120 and perform a function for adsorbing oil contained in the
refrigerant. The predetermined size may be about 1 .mu.m.
[0247] The third filter 330 may include a thread that is wound
around the gas inlet 123a. The thread may be made of a polyethylene
terephthalate (PET) material and have a predetermined thickness or
diameter.
[0248] A thickness or diameter of the thread may be determined to
have appropriate dimension in consideration of a strength of the
thread. When the thickness or diameter of the thread is too small,
the thread may be easily broken due to a very weak strength
thereof. On the other hand, if the thickness or diameter of the
thread is too large, a filtering effect with respect to foreign
matter may be deteriorated due to a very large pore in the gas
inlet 122 when the thread is wound. For example, the thickness or
diameter of the thread may be formed in hundreds of .mu.m units,
and the thread may be configured by combining several tens of .mu.m
units of spun threads into multiple strands.
[0249] The thread may be wound a number of times and be configured
such that an end thereof is fixed with a knot. The number of times
the thread is wound may be appropriately selected in consideration
of a degree of pressure drop of the gas refrigerant and a filtering
effect of foreign matter. When the number of windings is too large,
the pressure drop of the gas refrigerant becomes too large, and
when the number of windings is too small, foreign matter may not be
filtered well.
[0250] A tension force of the thread may be formed in an
appropriate size in consideration of a deformation degree of the
cylinder 120 and a fixing force of the thread. When the tension
force is too large, deformation of the cylinder 120 may be caused,
and when the tension force is too small, the thread may not be well
fixed to the gas inlet 122.
[0251] FIG. 17 is a cross-sectional view showing a state in which
the frame and the cylinder are coupled to each other according to
an embodiment. FIG. 18 is an enlarged view of portion "B" of FIG.
17.
[0252] Referring to FIGS. 17 and 18, the linear compressor 100
according to this embodiment may include a sealing pocket 370 that
communicates with the third passage 430 and in which a sealing
member 350 may be installed. The sealing pocket 370 may be a space
in which the sealing member 350 may be installed, and may be formed
between an inner circumferential surface of the frame body 111 and
the outer circumferential surface of the cylinder body 121. The
sealing pocket 370 may be formed at a rear portion of the frame 110
and the cylinder 120. Based on the flow direction of the
refrigerant, a flow cross-sectional area of the sealing pocket 370
may be formed to be larger than a flow cross-sectional area of the
third passage 430.
[0253] The rear portion of the frame body 111 may include a pocket
forming portion 112 recessed radially outward from the inner
circumferential surface of the frame body 111. The pocket forming
portion 112 may form at least one surface of the sealing pocket
370.
[0254] The frame main body 111 may further include a second
inclined portion 113 that extends obliquely from the pocket forming
portion 112 in a rearward inward direction. The cylinder body 121
may include a first inclined portion 128 that forms the sealing
pocket 370. The first inclined portion 128 may constitute at least
one surface of the sealing pocket 370.
[0255] The first inclined portion 128 may extend in an inclined
shape rearward from the first body end 121a of the cylinder body
121. The first inclined portion 128 may extend from an inside of
the pocket forming portion 112 to a point corresponding to an
inside of the second inclined portion 113.
[0256] Due to the recessed structure of the pocket forming portion
112 and the inclined structure of the first inclined portion 128, a
radial height of the sealing pocket 370 may be formed to be larger
than a diameter of the sealing member 350. An axial length of the
sealing pocket 370 may be larger than the diameter of the sealing
member 350. That is, the sealing pocket 370 may have a size such
that the sealing member 350 is movable without interfering with the
frame body 111 or the cylinder body 121.
[0257] An interval or distance between a rear portion of the first
inclined portion 128 and a rear portion of the second inclined
portion 113 may be smaller than the diameter of the sealing member
350. Therefore, when refrigerant flows backward along the third
passage 430 between operations of the linear compressor 100, the
sealing member 350 may move rearward due to the pressure of the
refrigerant and seal the spaced space. As such, as the sealing
member 350 may be disposed between the cylinder 120 and the frame
110 to seal the third passage 430, it is possible to prevent the
refrigerant in the third passage 430 from leaking outside of the
frame 110.
[0258] When the sealing member 350 is provided to be movable in the
sealing pocket 370 and the compressor is driven such that the
refrigerant in the third passage 430 flows, the sealing member 350
may be pressed against the cylinder 120 and the frame 110. Thus,
deformation of the cylinder 120 due to a pressing force of the
sealing member 350 may be prevented.
[0259] Hereinafter, flow of the refrigerant between operations of
the linear compressor will be described.
[0260] FIG. 19 is a cross-sectional view showing refrigerant flow
of a linear compressor according to an embodiment. FIG. 20 is a
view showing flow of refrigerant discharged from a compression
chamber in first and second passages according to an embodiment.
FIG. 21 is a view showing the flow of refrigerant in a third
passage.
[0261] First, the refrigerant flow in the linear compressor will be
described with reference to FIG. 19. Referring to FIG. 19, the
refrigerant flows into the inside of the shell 101 through the
suction inlet 104 and suction guide 155 to the suction muffler 150.
The refrigerant flows into the second muffler 153 through the first
muffler 151 of the suction muffler 150 and flows into the piston
130. In this process, suction noise of the refrigerant may be
reduced.
[0262] As the refrigerant passes through the first filter 310
provided in the suction muffler 150, foreign matter having a
predetermined size (25 .mu.m) or more may be filtered. The
refrigerant that passes through the suction muffler 150 and exists
inside of the piston 130 is suctioned into the compression space P
through the suction hole 133 when the suction valve 135 is
opened.
[0263] When the refrigerant pressure in the compression space P
exceeds the discharge pressure, the discharge valve 161 is opened
and the refrigerant is discharged to the discharge space of the
discharge cover 160 through the open discharge valve 161. The
discharge valve 161 is moved forward and spaced apart from the
front surface of the cylinder 120. In this process, the valve
spring 162 is elastically deformed forward. The stopper 163 limits
the amount of deformation of the valve spring 162 to a certain
extent. The refrigerant discharged into the discharge space 160a of
the discharge cover 160 flows to the discharge outlet 105 through
the loop pipe 165 coupled to the discharge cover 160 and is
discharged outside of the compressor 100.
[0264] At least a portion of the refrigerant in the discharge space
160a of the discharge cover 160 may flow through the space between
the cylinder 120 and the frame 110, that is, the first passage 410
and the second passage 420. The refrigerant may be filtered by the
second filter 320 in the process of flowing through the first
passage 410 or the second passage 420.
[0265] The filtered refrigerant may flow toward the outer
circumferential surface of the cylinder body 121 through the third
passage 430, and at least a portion of the refrigerant may flow
into the plurality of gas inlets 122 formed in the cylinder body
121. The refrigerant introduced into the gas inlet 122 may be
filtered by the third filter 330 and introduced into the cylinder
120 through the nozzle 123.
[0266] The refrigerant introduced into the cylinder 120 may flow
between the inner circumferential surface of the cylinder 120 and
the outer circumferential surface of the piston 130 and act to
separate the piston 130 from the inner circumferential surface of
the cylinder 120 (gas bearing). As such, the high pressure gas
refrigerant may be bypassed into the cylinder 120 to provide lift
pressure to the reciprocating piston 130 and may act as a bearing
at the interface between the cylinder and the piston. Therefore,
wear between the piston 130 and the cylinder 120 may be reduced. As
an oil bearing is not used, friction loss due to oil may not be
generated even when the compressor 100 is operated at high
speed.
[0267] The refrigerant flowing to the inner surface of the cylinder
and the outer surface of the piston may be a compressed refrigerant
and have a condensing pressure (Pd). In contrast, the space in
which the refrigerant flows has an evaporating pressure (Ps) before
being compressed. Therefore, as the difference between the
condensing pressure and the evaporation pressure is greater, the
piston 130 may be lifted from the inner surface of the cylinder 120
with a greater force.
[0268] In addition, as a plurality of filters is provided in the
passage of the refrigerant flowing inside of the compressor 100, it
is possible to remove foreign matter contained in the refrigerant.
This may improve reliability of the refrigerant to act as the gas
bearing. Therefore, it is possible to prevent a phenomenon in which
abrasion is generated in the piston 130 or the cylinder 120 by
foreign matter contained in the refrigerant.
[0269] As oil contained in the refrigerant is removed by the
plurality of filters, it is possible to prevent friction loss
caused by oil. The first filter 310, the second filter 320, and the
third filter 330 may be collectively referred to as a "refrigerant
filter device" in that the first filter 310, the second filter 320,
and the third filter 330 filter the refrigerant to act as the gas
bearing.
[0270] The refrigerant flowing through the third passage 430 acts
on the sealing member 350. That is, the pressure of the refrigerant
acts on the sealing member 350, and the sealing member 350 is moved
from the sealing pocket 370 to a point between the first inclined
portion 128 of the cylinder 120 and the second inclined portion 113
of the frame 110.
[0271] The sealing member 350 is in close contact with the cylinder
120 and the frame 110 and seals the space between the cylinder 120
and the frame 110, for example, the space between the first
inclined portion 128 and the second inclined portion 113.
Therefore, it is possible to prevent refrigerant in the third
passage 430 from leaking to the outside through the space between
the cylinder 120 and the frame 110.
[0272] When driving of the linear compressor 100 is stopped, the
pressure of the refrigerant acting on the sealing member 350 is
released. Thus, the contact force between the sealing member 350,
the cylinder 120, and the frame 110 is weakened. As a result, the
sealing member 350 is freely movable within the sealing pocket 220,
for example, in a state of being spaced apart from the first
inclined portion 128 and the second inclined portion 113 (indicated
by a dashed line).
[0273] According to this action, the sealing member 350 is in close
contact with the cylinder 120 and the frame 110 only when the
compressor 100 is driven, and thus, sealing of the third passage
430 may be performed. Therefore, the force applied to the cylinder
120 from the sealing member 350 may be reduced. Therefore,
deformation of the cylinder 120 may be prevented.
[0274] As the sealing member 350 may be placed in a movable state
in the sealing pocket 370, it is possible to prevent interference
of the sealing member 350 when the cylinder 120 and the frame 110
are assembled. As a result, assembling of the cylinder 120 and the
frame 110 may be facilitated.
[0275] As described above, as the non-azeotropic mixed refrigerant
is used, the lift pressure of the piston may be increased, and
operation of the gas bearing may be performed smoothly. As a
minimum lift pressure for lift of the piston may be provided even
in a low cooling power operation period, reliability of the linear
compressor may be improved. For example, as the lift pressure is
increased and friction at the contact surface between the cylinder
and the piston decreases, it can advantageously act to increase
efficiency of the linear compressor through mechanical design, such
as the gas discharge hole and the filter. It is possible to
maximize the effect by applying the non-azeotropic mixed
refrigerant to the oilless linear compressor in the refrigerating
apparatus that performs the continuous operation mode in which the
low cooling power operation is frequently performed.
[0276] In comparison with a linear compressor using oil, when
performing long time low cooling power driving for the continuous
operation mode, the oilless linear compressor has an advantage in
that the oil is heated by friction between the cylinder and the
piston to reduce evaporation and combustion. In the above
description, the linear compressor is employed as the compressor.
However, embodiments are not limited thereto and may be applied to
all compressors in which oil circulation and air bearing are
performed due to the pressure difference of the refrigerant during
compression of the gas refrigerant.
INDUSTRIAL APPLICABILITY
[0277] According to embodiments disclosed herein, as a
non-azeotropic mixed refrigerant is used, it is possible to improve
durability by preventing wear of components due to a low
evaporation temperature of a refrigerant in a continuous operation
mode. In addition, it is possible to smoothly operate oil
circulation and air bearings for lubrication essential to operation
of a compressor.
* * * * *