U.S. patent application number 09/725537 was filed with the patent office on 2001-05-31 for control valve in variable displacement compressor.
Invention is credited to Kawaguchi, Masahiro, Kimura, Kazuya, Ota, Masaki, Suitou, Ken.
Application Number | 20010002237 09/725537 |
Document ID | / |
Family ID | 26576699 |
Filed Date | 2001-05-31 |
United States Patent
Application |
20010002237 |
Kind Code |
A1 |
Ota, Masaki ; et
al. |
May 31, 2001 |
Control valve in variable displacement compressor
Abstract
A control valve for a variable displacement compressor in
refrigerant circuit permits the compressor displacement to be
accurately controlled regardless of the thermal load on an
evaporator. The refrigerant circuit includes a high pressure pipe,
which extends between a discharge chamber of the compressor and a
condenser. A first pressure monitoring point is located in the
discharge chamber. A second pressure monitoring point is located in
the high pressure pipe. A supply passage connects the second
pressure monitoring point with a crank chamber of the compressor.
The control valve is located in the supply passage and adjusts the
opening size of the supply passage in accordance with the
difference between the pressure at the first pressure monitoring
point and the pressure at the second pressure monitoring point. The
control valve includes a solenoid for determining the target value
of the pressure difference. The control valve operates to maintain
the determined target value.
Inventors: |
Ota, Masaki; (Kariya-shi,
JP) ; Kimura, Kazuya; (Kariya-shi, JP) ;
Kawaguchi, Masahiro; (Kariya-shi, JP) ; Suitou,
Ken; (Kariya-shi, JP) |
Correspondence
Address: |
Kurt E. Richter
Morgan & Finnegan, L.L.P.
345 Park Avenue
New York
NY
10154
US
|
Family ID: |
26576699 |
Appl. No.: |
09/725537 |
Filed: |
November 29, 2000 |
Current U.S.
Class: |
417/222.2 ;
251/129.02; 251/129.15 |
Current CPC
Class: |
F04B 2027/1813 20130101;
F04B 27/1804 20130101; F04B 2027/1859 20130101; F04B 2207/03
20130101; F04B 2027/185 20130101; F04B 2027/1827 20130101; F04B
2205/08 20130101 |
Class at
Publication: |
417/222.2 ;
251/129.15; 251/129.02 |
International
Class: |
F04B 001/26 |
Foreign Application Data
Date |
Code |
Application Number |
Nov 30, 1999 |
JP |
11-340401 |
Mar 17, 2000 |
JP |
2000-075538 |
Claims
What is claimed is:
1. A control valve for a variable displacement compressor used in a
refrigerant circuit, wherein the refrigerant circuit includes a
condenser and a high pressure passage extending from a discharge
chamber of the compressor to the condenser, wherein a section of
the refrigerant circuit that includes the discharge chamber, the
condenser and the high pressure passage forms a high pressure zone,
and wherein the control valve controls the pressure in a crank
chamber of the compressor to change the displacement of the
compressor, the control valve comprising: a valve housing, wherein
the valve housing is located in a supply passage, which connects
the high pressure zone to the crank chamber, wherein the supply
passage includes an upstream section, which is between the high
pressure zone and the valve housing, and a downstream section,
which is between the valve housing and the crank chamber; a first
pressure chamber defined in the valve housing, the first pressure
chamber being exposed to the pressure of a first pressure
monitoring point, which is located in the high pressure zone; a
second pressure chamber defined in the valve housing, the second
pressure chamber being exposed to the pressure of a second pressure
monitoring point, which is located in a part of the high pressure
zone that is downstream of the first pressure monitoring point,
wherein the upstream section of the supply passage connects the
first pressure chamber or the second pressure chamber to the
corresponding pressure monitoring point; a valve body located in
the valve housing, wherein the valve body adjusts the opening size
of the supply passage; and a pressure receiving body located in the
valve housing, wherein the pressure receiving body moves the valve
body in accordance with the difference between the pressure in the
first pressure chamber and the pressure in the second pressure
chamber.
2. The control valve according to claim 1, wherein the pressure
receiving body is a rod, which moves axially, and wherein the rod
has an end face that receives the pressure of the first pressure
chamber and another end face that receives the pressure in the
second pressure chamber.
3. The control valve according to claim 2, wherein the valve body
is integral with the rod.
4. The control valve according to claim 1, wherein a valve chamber
for accommodating the valve body and a through hole for
communicating the valve chamber with the first pressure chamber are
defined in the valve housing, wherein the pressure receiving body
includes a divider and a coupler, wherein the divider is located in
the through hole to disconnect the valve chamber from the first
pressure chamber and the coupler couples the divider with the valve
body, and wherein the cross-sectional area of the coupler is less
than the cross-sectional area of the through hole.
5. The control valve according to claim 4, wherein the
cross-sectional area of the divider is equal to the cross-sectional
area of a section of the through hole that opens to the valve
chamber.
6. The control valve according to claim 1, further comprising an
actuator for urging the valve body by a force, the magnitude of
which corresponds to an external signal, wherein the urging force
of the actuator represents the target value of the pressure
difference, and wherein the pressure receiving body moves the valve
body such that the pressure difference seeks the target value.
7. The control valve according to claim 6, wherein the actuator
urges the valve body in a direction opposite to the direction of
the force applied to the pressure receiving body based on the
pressure difference.
8. The control valve according to claim 6, wherein the actuator is
a solenoid that generates an electromagnetic force, the magnitude
of which corresponds to the magnitude of a supplied current,
wherein the control valve includes an urging member that urges the
valve body in a direction opposite to the direction in which the
solenoid urges the valve body, and wherein, when electric current
is not supplied to the solenoid, the urging member causes the valve
body to maximize the opening size of the supply passage.
9. The control valve according to claim 6, wherein the actuator
includes a plunger chamber and a plunger accommodated in the
plunger chamber, the plunger chamber functioning as either the
first pressure chamber or the second pressure chamber, wherein the
pressure receiving body is a rod, which moves axially, and wherein
the rod includes an end that extends into the plunger chamber and
is fixed to the plunger.
10. The control valve according to claim 9, wherein the end of the
rod that is fixed to the plunger is a first end, and wherein the
rod includes a second end that extends into the pressure chamber
other than the plunger chamber.
11. A displacement control mechanism for a variable displacement
compressor used in a refrigerant circuit, wherein the refrigerant
circuit includes a condenser and a high pressure passage extending
from a discharge chamber of the compressor to the condenser,
wherein a section of the refrigerant circuit that includes the
discharge chamber, the condenser and the high pressure passage
forms a high pressure zone, and wherein the displacement control
mechanism controls the pressure in a crank chamber of the
compressor to change the displacement of the compressor, the
displacement control mechanism comprising: a supply passage,
wherein the supply passage connects the high pressure zone to the
crank chamber to conduct gas from the high pressure zone to the
crank chamber; a control valve located in the supply passage,
wherein the control valve includes a first pressure chamber and a
second pressure chamber, and wherein the supply passage includes an
upstream section, which is between the high pressure zone and the
control valve, and a downstream section, which is between the
control valve and the crank chamber; a first introduction passage
for connecting a first pressure monitoring point, which is located
in the high pressure zone, with the first pressure chamber; a
second introduction passage for connecting a second pressure
monitoring point, which is located in a part of the high pressure
zone that is downstream of the first pressure monitoring point,
with the second pressure chamber, wherein either the first
introduction passage or the second introduction passage functions
as the upstream section of the supply passage; a valve body located
in the control valve, wherein the valve body adjusts the opening
size of the supply passage; and a pressure receiving body located
in the control valve, wherein the pressure receiving body moves the
valve body in accordance with the difference between the pressure
in the first pressure chamber and the pressure in the second
pressure chamber.
12. The displacement control mechanism according to claim 11,
wherein the control valve includes a valve chamber for
accommodating the valve body and a through hole for communicating
the valve chamber with the first pressure chamber, wherein the
pressure receiving body includes a divider and a coupler, wherein
the divider is located in the through hole to disconnect the valve
chamber from the first pressure chamber and the coupler couples the
divider with the valve body, and wherein the cross-sectional area
of the coupler is less than the cross-sectional area of the through
hole.
13. The displacement control mechanism according to claim 12,
wherein the cross-sectional area of the divider is equal to the
cross-sectional area of a section of the through hole that opens to
the valve chamber.
14. The displacement control mechanism according to claim 11,
wherein the control valve further comprises an actuator for urging
the valve body by a force, the magnitude of which corresponds to an
external signal, wherein the urging force of the actuator
represents the target value of the pressure difference, and wherein
the pressure receiving body moves the valve body such that the
pressure difference seeks the target value.
15. The displacement control mechanism according to claim 14,
wherein the actuator urges the valve body in a direction opposite
to the direction of the force applied to the pressure receiving
body based on the pressure difference.
16. The displacement control mechanism according to claim 14,
wherein the actuator is a solenoid that generates an
electromagnetic force, the magnitude of which corresponds to the
magnitude of a supplied current, wherein the control valve includes
an urging member that urges the valve body in a direction opposite
to the direction in which the solenoid urges the valve body, and
wherein, when electric current is not supplied to the solenoid, the
urging member causes the valve body to maximize the opening size of
the supply passage.
17. The displacement control mechanism according to claim 14,
wherein the actuator includes a plunger chamber and a plunger
accommodated in the plunger chamber, the plunger chamber
functioning as either the first pressure chamber or the second
pressure chamber, wherein the pressure receiving body is a rod,
which moves axially, and wherein the rod includes an end that
extends into the plunger chamber and is fixed to the plunger.
18. The displacement control mechanism according to claim 17,
wherein the end of the rod that is fixed to the plunger is a first
end, and wherein the rod includes a second end that extends into
the pressure chamber other than the plunger chamber.
19. The displacement control mechanism according to claim 11,
wherein a fixed restrictor is located in the high pressure passage
between the first pressure monitoring point and the second pressure
monitoring point.
Description
BACKGROUND OF THE INVENTION
[0001] The present invention relates to a control Valve used in a
variable displacement compressor. More particularly, the present
invention pertains to a control valve that controls the compressor
displacement by adjusting the pressure in a crank chamber.
[0002] A typical refrigerant circuit of a vehicle air conditioner
includes a condenser, an expansion valve, an evaporator and a
compressor. The compressor receives refrigerant gas from the
evaporator. The compressor then compresses the gas and discharges
the gas to the condenser. The evaporator transfers heat to the
refrigerant in the refrigerant circuit from the air in the
passenger compartment. The pressure of refrigerant gas at the
outlet of the evaporator, in other words, the pressure of
refrigerant gas that is drawn into the compressor (suction pressure
Ps), represents the thermal load on the refrigerant circuit.
[0003] Variable displacement swash plate type compressors are
widely used in vehicles. Such compressors include a displacement
control valve that operates to maintain the suction pressure Ps at
a predetermined target level (target suction pressure). The control
valve changes the inclination angle of the swash plate in
accordance with the suction pressure Ps for controlling the
displacement of the compressor. The control valve includes a valve
body and a pressure sensing member such as a bellows or a
diaphragm. The pressure sensing member moves the valve body in
accordance with the suction pressure Ps, which adjusts the pressure
in a crank chamber. The inclination of the swash plate is adjusted,
accordingly.
[0004] In addition to the above structure, some control valves
include an electromagnetic actuator, such as a solenoid, to change
the target suction pressure. An electromagnetic actuator urges a
pressure sensing member or a valve body in one direction by a force
that corresponds to the value of an externally supplied current.
The magnitude of the force determines the target suction pressure.
Varying the target suction pressure permits the air conditioning to
be finely controlled.
[0005] Such compressors are usually driven by vehicle engines.
Among the auxiliary devices of a vehicle, the compressor consumes
the most engine power and is therefore a great load on the engine.
When the load on the engine is great, for example, when the vehicle
is accelerating or moving uphill, all available engine power needs
to be used for moving the vehicle. Under such conditions, to reduce
the engine load, the compressor displacement is minimized. This
will be referred to as a displacement limiting control procedure. A
compressor having a control valve that changes a target suction
pressure raises the target suction pressure when executing the
displacement limiting control procedure. Then, the compressor
displacement is decreased such that the actual suction pressure Ps
is increased to approach the target suction pressure.
[0006] The graph of FIG. 11 illustrates the relationship between
suction pressure Ps and displacement Vc of a compressor. The
relationship is represented by multiple lines in accordance with
the thermal load in an evaporator. Thus, if the suction pressure Ps
is constant, the compressor displacement Vc increases as the
thermal load increases. If a level Ps1 is set as a target suction
pressure, the actual displacement Vc varies in a certain range
(.DELTA.Vc in FIG. 11) due to the thermal load. If a high thermal
load is applied to the evaporator during the displacement limiting
control procedure, an increase of the target suction pressure does
not lower the compressor displacement Vc to a level that
sufficiently reduces the engine load.
[0007] Thus, the compressor displacement is not always controlled
as desired as long as the displacement is controlled based on the
suction pressure Ps.
SUMMARY OF THE INVENTION
[0008] Accordingly, it is an objective of the present invention to
provide a control valve used in a variable displacement compressor
that accurately controls the compressor displacement regardless of
the thermal load on an evaporator.
[0009] To achieve the above objective, the present invention
provides a control valve for a variable displacement compressor
used in a refrigerant circuit. The refrigerant circuit includes a
condenser and a high pressure passage extending from a discharge
chamber of the compressor to the condenser. A section of the
refrigerant circuit that includes the discharge chamber, the
condenser and the high pressure passage forms a high pressure zone.
The control valve controls the pressure in a crank chamber of the
compressor to change the displacement of the compressor. The
control valve includes a valve housing. The valve housing is
located in a supply passage, which connects the high pressure zone
to the crank chamber, The supply passage includes an upstream
section, which is between the high pressure zone and the valve
housing, and a downstream section, which is between the valve
housing and the crank chamber. A first pressure chamber is defined
in the valve housing. The first pressure chamber is exposed to the
pressure of a first pressure monitoring point, which is located in
the high pressure zone. A second pressure chamber is defined in the
valve housing. The second pressure chamber is exposed to the
pressure of a second pressure monitoring point, which is located in
a part of the high pressure zone that is downstream of the first
pressure monitoring point. The upstream section of the supply
passage connects the first pressure chamber or the second pressure
chamber to the corresponding pressure monitoring point. A valve
body is located in the valve housing. The valve body adjusts the
opening size of the supply passage. A pressure receiving body is
located in the valve housing. The pressure receiving body moves the
valve body in accordance with the difference between the pressure
in the first pressure chamber and the pressure in the second
pressure chamber.
[0010] Other aspects and advantages of the invention will become
apparent from the following description, taken in conjunction with
the accompanying drawings, illustrating by way of example the
principles of the invention.
BRIEF DESCRIPTION OF THE DRAWINGS
[0011] The invention, together with objects and advantages thereof,
may best be understood by reference to the following description of
the presently preferred embodiments together with the accompanying
drawings in which:
[0012] FIG. 1 is a cross-sectional view illustrating a variable
displacement swash plate type compressor according to a first
embodiment of the present invention;
[0013] FIG. 2 is a schematic diagram illustrating a refrigerant
circuit including the compressor of FIG. 1;
[0014] FIG. 3 is a cross-sectional view illustrating a control
valve of FIG. 1;
[0015] FIG. 4 is a schematic cross-sectional view showing part of
the control valve shown in FIG. 3;
[0016] FIG. 5 is a flowchart showing a main routine for controlling
a compressor displacement;
[0017] FIG. 6 is a flowchart showing a normal control
procedure;
[0018] FIG. 7 is a flow chart showing an exceptional control
procedure;
[0019] FIG. 8 is a cross-sectional view illustrating a control
valve according to a second embodiment of the present
invention;
[0020] FIG. 9 is a cross-sectional view illustrating a control
valve according to a third embodiment of the present invention;
[0021] FIG. 10 is a cross-sectional view showing part of a control
valve according to a fourth embodiment of the present invention;
and
[0022] FIG. 11 is a graph showing the relationship between the
suction pressure Ps and the displacement Vc of a prior art
compressor.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0023] A first embodiment of the present invention will now be
described with reference to FIGS. 1 to 7. As shown in FIG. 1, a
variable displacement swash plate type compressor used in a vehicle
includes a cylinder block 11, a front housing member 12, which is
secured to the front end face of the cylinder block 11, and a rear
housing member 14, which is secured to the rear end face of the
cylinder block 11. A valve plate assembly 13 is located between the
cylinder block 11 and the rear housing member 14. In FIG. 1, the
left end of the compressor is defined as the front end, and the
right end of the compressor is defined as the rear end.
[0024] A crank chamber 15 is defined between the cylinder block 11
and the front housing member 12. A drive shaft 16 extends through
the crank chamber 15 and is supported by the cylinder block 11 and
a front housing member 12.
[0025] The front end of the drive shaft 16 is connected to an
external drive source, which is an engine Eg in this embodiment,
through a power transmission mechanism PT. The power transmission
mechanism PT includes a belt and a pulley. The mechanism PT may be
a clutch mechanism, such as an electromagnetic clutch, which is
electrically controlled from the outside. In this embodiment, the
mechanism PT has no clutch mechanism. Thus, when the engine Eg is
running, the compressor is driven continuously.
[0026] A lug plate 17 is secured to the drive shaft 16 in the crank
chamber 15. A drive plate, which is a swash plate 18 in this
embodiment, is accommodated in the crank chamber 15. The swash
plate 18 has a hole formed in the center. The drive shaft 16
extends through the hole in the swash plate 18. The swash plate 18
is coupled to the lug plate 17 by a hinge mechanism 19. The hinge
mechanism 19 permits the swash plate 18 to rotate integrally with
the lug plate 17 and drive shaft 16. The hinge mechanism 19 also
permits the swash plate 18 to slide along the drive shaft 16 and to
tilt with respect to a plane perpendicular to the axis of the drive
shaft 16.
[0027] Several cylinder bores 20 (only one shown) are formed about
the axis of the drive shaft 16 in the cylinder block 11. A single
headed piston 21 is accommodated in each cylinder bore 20. Each
piston 21 and the corresponding cylinder bore 20 define a
compression chamber. Each piston 21 is coupled to the swash plate
18 by a pair of shoes 28. The swash plate 18 coverts rotation of
the drive shaft 16 into reciprocation of each piston 21.
[0028] A suction chamber 22 and a discharge chamber 23 are defined
between the valve plate assembly 13 and the rear housing member 14.
The suction chamber 22 forms a suction pressure zone, the pressure
of which is a suction pressure Ps. The discharge chamber 23 forms a
discharge pressure zone, the pressure of which is a discharge
pressure Pd. The valve plate assembly 13 has suction ports 24,
suction valve flaps 25, discharge ports 26 and discharge valve
flaps 27. Each set of the suction port 24, the suction valve flap
25, the discharge port 26 and the discharge valve flap 27
corresponds to one of the cylinder bores 20. When each piston 21
moves from the top dead center position to the bottom dead center
position, refrigerant gas in the suction chamber 22 flows into the
corresponding cylinder bore 20 via the corresponding suction port
24 and suction valve 25. When each piston 21 moves from the bottom
dead center position to the top dead center position, refrigerant
gas in the corresponding cylinder bore 20 is compressed to a
predetermined pressure and is discharged to the discharge chamber
23 via the corresponding discharge port 26 and discharge valve
27.
[0029] The inclination angle of the swash plate 18 is determined
according to the pressure in the crank chamber 15 (crank pressure
Pc). The inclination angle of the swash plate 18 defines the stroke
of each piston 21 and the displacement of the compressor.
[0030] As shown in FIGS. 1 and 2, the refrigerant circuit of the
vehicle air conditioner includes the compressor and an external
circuit 35, which is connected to the compressor. The external
circuit 35 includes a condenser 36, a temperature-type expansion
valve 37 and an evaporator 38. The expansion valve 37 adjusts the
flow rate of refrigerant supplied to the evaporator 38 based on the
temperature or the pressure detected by a heat sensitive tube 37a,
which is located downstream of the evaporator 38. The temperature
or the pressure at the downstream of the evaporator 38 represents
the thermal load on the evaporator 38. The external circuit 35
includes a low pressure pipe 39, which extends from the evaporator
38 to the suction chamber 22 of the compressor, and a high pressure
pipe 40, which extends from the discharge chamber 23 of the
compressor to the condenser 36.
[0031] The flow rate of the refrigerant in the refrigerant circuit
is expressed by the product of the amount of the refrigerant gas
discharged from the compressor during one rotation of the drive
shaft 16 multiplied by the rotational speed of the drive shaft 16.
Under the condition where the engine Eg rotates at a constant
rotational speed, the flow rate of the refrigerant in the
refrigerant circuit increases as the compressor displacement
increases when the inclination angle of the swash plate 18
increases. In other words, when the inclination angle of the swash
plate 18 or the compressor displacement is constant, the flow rate
of the refrigerant in the refrigerant circuit increases as the
rotational speed of the engine Eg increases.
[0032] Pressure loss in the refrigerant circuit increases as the
flow rate of the refrigerant in the refrigerant circuit increases.
If an upstream first pressure monitoring point and a downstream
second pressure monitoring point are set up in the refrigerant
circuit, the pressure difference between these two points due to
the pressure loss shows a positive correlation with the flow rate
of the refrigerant in the refrigerant circuit. Thus, the flow rate
of the refrigerant in the refrigerant circuit can be detected
indirectly by detecting the difference between the refrigerant gas
pressure at the first pressure monitoring point and that at the
second pressure monitoring point. In this embodiment, a first
pressure monitoring point P1 is set up in the discharge chamber 23
corresponding to the most upstream section in the high pressure
pipe 40, and a second pressure monitoring point P2 is set up in the
high pressure pipe 40 at a predetermined distance downstream from
the first point P1, as shown in FIG. 2. The refrigerant gas
pressure at the first pressure monitoring point P1 and that at the
second pressure monitoring point P2 are hereinafter referred to as
PdH and PdL, respectively.
[0033] The compressor has a crank pressure control mechanism for
controlling the crank pressure Pc. As shown in FIGS. 1 and 2, the
crank pressure control mechanism includes a bleed passage 31, a
first pressure introduction passage 41, a second pressure
introduction passage 42, a crank passage 44 and a control valve
100. The bleed passage 31 connects the crank chamber 15 to the
suction chamber 22 to conduct refrigerant gas from the crank
chamber 15 to the suction chamber 22. The first pressure
introduction passage 41 connects the discharge chamber 23, i.e.,
the first pressure monitoring point P1, to the control valve 100.
The second pressure introduction passage 42 connects the second
pressure monitoring point P2 to the control valve 100. The crank
passage 44 connects the control valve 100 to the crank chamber
15.
[0034] The second pressure introduction passage 42 and the crank
passage 44 form a supply passage 110 for connecting the second
pressure monitoring point P2 to the crank chamber 15. The second
pressure introduction passage 42 forms an upstream section of the
supply passage 110, and the crank passage 44 forms a downstream
section of the supply passage 110. The control valve 100 adjusts
the flow rate of the high pressure refrigerant gas supplied from
the second pressure monitoring point P2, through the supply passage
110, to the crank chamber 15 to control the crank pressure Pc.
[0035] As shown in FIG. 2, the high pressure pipe 40 is provided
with a fixed restrictor 43 between the first pressure monitoring
point P1 and the second pressure monitoring point P2. The fixed
restrictor 43 increases the pressure difference (PdH-PdL) between
the two pressure monitoring points P1 and P2. This enables the
distance between the two pressure monitoring points P1 and P2 to be
reduced and permits the second pressure monitoring point P2 to be
relatively close to the compressor. Thus, the second pressure
introduction passage 42, which extends from the second pressure
monitoring point P2 to the control valve 100 in the compressor, can
be shortened.
[0036] As shown in FIG. 1, the control valve 100 is fitted in a
receiving hole 14a of the rear housing member 14. As shown in FIGS.
3 and 4, the control valve 100 is provided with an inlet valve
mechanism 51 and a solenoid 52, which serves as an electromagnetic
actuator. The inlet valve mechanism 51 adjusts the aperture of the
supply passage 110. The solenoid 52 exerts a force according to the
level of the electric current supplied from the outside to the
inlet valve mechanism 51 through an operating rod 53. The operating
rod 53 is cylindrical and has a divider 54, a coupler 55 and a
guide 57. The part of the guide 57 adjacent to the coupler 55
functions as a valve body 56. The cross-sectional area S3 of the
coupler 55 is smaller than the cross-sectional area S4 of the guide
57 and the valve body 56.
[0037] The control valve 100 has a valve housing 58 containing an
upper housing member 58b and a lower housing member 58c. The upper
housing member 58b constitutes a shell for the inlet valve
mechanism 51, and the lower housing member 58c constitutes a shell
for the solenoid 52. A plug 58a is screwed into the upper housing
member 58b to close an opening in its upper end. A valve chamber 59
and a through hole 60 connected thereto are defined in the upper
housing member 58b. The upper housing member 58b and the plug 58a
define a high pressure chamber 65 as a first pressure chamber. The
high pressure chamber 65 and the valve chamber 59 communicate with
each other through the through hole 60. The operating rod 53
extends through the valve chamber 59, the through hole 60 and the
high pressure chamber 65. The operating rod 53 moves axially such
that the valve body 56 selectively connects and blocks off the
valve chamber 59 with respect to the through hole 60.
[0038] A first radial port 62 is formed in the upper housing member
58b to communicate with the valve chamber 56. The valve chamber 59
is connected to the second pressure monitoring point P2 through the
first port 62 and the second pressure introduction passage 42.
Thus, the pressure PdL at the second pressure monitoring point P2
exerts to the inside of the valve chamber 59 through the second
pressure introduction passage 42 and the first port 62. A second
port 63 extending radially is formed in the upper housing member
58b to communicate with the through hole 60. The through hole 60 is
connected to the crank chamber 15 through the second port 63 and
the crank passage 44. When the valve body 56 opens to connect the
valve chamber 59 to the through hole 60, the refrigerant gas is
supplied from the second pressure monitoring point P2, through the
supply passage 110, which includes the second pressure introduction
passage 42 and the crank passage 44, into the crank chamber 15. The
ports 62 and 63, the valve chamber 59 and the through hole 60
constitute a part of the supply passage 110 within the control
valve 100.
[0039] The valve body 56 is located in the valve chamber 59. The
cross-sectional area S3 of the coupler 55 is less than the
cross-sectional area S1 of the through hole 60. The cross-sectional
area S1 of the through hole 60 is less than the cross-sectional
area S4 of the valve body 56. The inner wall of the valve chamber
59, to which the through hole 60 opens, functions as a valve seat
64 for receiving the valve body 56. The through hole 60 functions
as a valve opening, which is opened and closed selectively by the
valve body 56. When the valve body 56 is abutted against the valve
seat 64, the through hole 60 is shut off from the valve chamber 59.
As shown in FIG. 3, when the valve body 56 is spaced from the valve
seat 64, the through hole 60 is connected to the valve chamber
59.
[0040] The divider 54 of the operating rod 53 has a portion located
in the through hole 60 and a portion located in the high pressure
chamber 65. The cross-sectional area S2 of the divider 54 is equal
to the cross-sectional area S1 of the through hole 60. Therefore,
the divider 54 shuts off the high pressure chamber 65 from the
valve chamber 59.
[0041] A third radial port 67 is defined in the upper housing
member 58b to communicate with the high pressure chamber 65. The
high pressure chamber 65 is connected through the third port 67 and
the first pressure introduction passage 41 to the first pressure
monitoring point P1 or the discharge chamber 23. Thus, the pressure
PdH at the first pressure monitoring point P1 is exerted through
the first pressure introduction passage 41 and the third port 67 to
the high pressure chamber 65.
[0042] A return spring 68 is contained in the high pressure chamber
65. The return spring 68 urges the operating rod 53 to cause the
valve body 56 to move away from the valve seat 64 through an
aligning mechanism. The upper end of the return spring 68 is
received by the plug 58a. The position of the plug 58a can be
changed axially with respect to the upper housing member 58b. The
urging force of the return spring 68 is varied depending on the
axial position of the plug 58a with respect to the upper housing
member 58b.
[0043] The aligning mechanism contains a spring seat 79 for
receiving the return spring 68, and an aligning ball 80 located
between the valve seat 79 and the divider 54. The spring seat 79
and the divider 54 each have a conical recess in which the aligning
ball 80 is retained. The aligning mechanism corrects the action of
the return spring 68 such that the force of the return spring 68 is
applied in the axial direction. Even if the return spring 68 is
tilted with respect to the axial line of the operating rod 53, only
an axial force is applied to the operating rod 53. This provides
smooth and accurate operation of the operating rod 53.
[0044] A first seal ring 76 is fitted on the outer surface of the
lower housing member 58c. A second seal ring 77 and a third seal
ring 78 are fitted on the outer surface of the upper housing member
58b. When the control valve 100 is fitted in the receiving hole 14a
of the rear housing member 14 (see FIG. 1), the first, second and
third seal rings 76, 77, 78 contact the inner circumference of the
receiving hole 14a. The first seal ring 76 isolates the first port
62 from the outside of the compressor. The second seal ring 77
isolates the second port 63 from the first port 62. The third seal
ring 78 isolates the third port 67 from the second port 63.
[0045] The solenoid 52 is provided with a cup-shaped receiving
cylinder 69, which is fixed in the lower housing member 58c. A
fixed iron core 70 is fitted in the upper opening of the receiving
cylinder 69. The fixed iron core 70 constitutes a part of the inner
wall of the valve chamber 59 and also defines a plunger chamber 71,
which serves as a second pressure chamber. A plunger 72 is located
in this plunger chamber 71. The fixed iron core 70 includes a guide
hole 73, which accommodates the guide 57 of the operating rod 53. A
slight clearance (not shown) exists between the inner wall of the
guide hole 73 and the guide 57. The valve chamber 59 and the
plunger chamber 71 communicate normally with each other through the
clearance. Thus, the pressure in the valve chamber 59, or the
pressure PdL at the second pressure monitoring point P2, is applied
inside the plunger chamber 71.
[0046] The lower end of the guide 57 extends into the plunger
chamber 71. The plunger 72 is fixed to the lower end of the guide
57. The plunger 72 moves in the axial direction integrally with the
operating rod 53. A shock absorbing spring 74 is contained in the
plunger chamber 71 to urge the plunger 72 toward the fixed iron
core 70.
[0047] A coil 75 surrounds the fixed iron core 70 and the plunger
72. A controller 81 supplies electric power to the coil 75 through
a drive circuit 82. The coil 75 then generates an electromagnetic
force F between the fixed iron core 70 and the plunger 72
corresponding to the level of the electric power supplied to the
coil 75. The electromagnetic force F attracts the plunger 72 toward
the fixed iron core 70 and urges the operating rod 53 to cause the
valve body 56 to move toward the valve seat 64.
[0048] The force of the shock absorbing spring 74 is smaller than
the force of the return spring 68. Therefore, the return spring 68
moves the plunger 72 and the operating rod 53 to the initial
position as shown in FIG. 3 when no power is supplied to the coil
75, and the valve body 56 is moved to the lowest position to
maximize the opening size of the through hole 60.
[0049] There are methods for changing voltage applied to the coil
75, one of which is to change the voltage value and another is
referred to as PWM control or duty control. Duty control is
employed in this embodiment. Duty control is a method where the
ON-time per cycle of a pulsed voltage, which is turned on and off
periodically, is adjusted to modify the average value of the
voltage applied. An average applied voltage value can be obtained
by multiplying the value obtained by dividing the ON-time of the
pulsed voltage by the cycle time thereof, i.e., the duty ratio Dt,
by the pulsed voltage value. In duty control, the electric current
varies intermittently. This reduces hysteresis of the solenoid 52.
The smaller the duty ratio Dt is, the smaller the electromagnetic
force F generated between the fixed iron core 70 and the plunger 72
is and the greater the opening size of the through hole 60 by the
valve body 56 is. It is also possible to measure the value of the
electric current flowing through the coil 75 and perform feed back
control of the value of the voltage applied to the coil 75.
[0050] The opening size of the through hole 60 by the valve body 56
depends on the axial position of the operating rod 53. The axial
position of the operating rod 53 is determined based on various
forces that act axially on the operating rod 53. These forces will
be described referring to FIGS. 3 and 4. The downward forces in
FIGS. 3 and 4 tend to space the valve body 56 from the valve seat
64 (the valve opening direction). The upward forces in FIGS. 3 and
4 tend to move the valve body 56 toward the valve seat 64 (the
valve closing direction).
[0051] First, the various forces acting on the portion of the
operating rod 53 above the coupler 55, i.e., on the divider 54,
will be described. As shown in FIGS. 3 and 4, the divider 54
receives a downward force f1 from the return spring 68. The divider
54 also receives a downward force based on the pressure PdH in the
high pressure chamber 65. The effective pressure receiving area of
the divider 54 with respect to the pressure PdH in the high
pressure chamber 65 is equal to the cross-sectional area S2 of the
divider 54. The divider 54 also receives an upward force based on
the pressure in the through hole 60 (crank pressure Pc). The
effective pressure receiving area of the divider 54 with respect to
the pressure in the through hole 60 is equal to the cross-sectional
area S2 of the divider 54 minus the cross-sectional area S3 of the
coupler 55. Provided that the downward forces are positive values,
the net force .SIGMA.F1 acting upon the divider 54 can be expressed
by the following equation I.
.SIGMA.F1=PdH.multidot.S2-Pc(S2-S3)+f1 Equation I
[0052] Next, various forces that act upon the portion of the
operating rod 53 below the coupler 55, i.e., on the guide 57, will
be described. The guide 57 receives an upward force f2 from the
shock absorbing spring 74 and an upward electromagnetic force F
from the plunger 72. Further, as shown in FIG. 4, the end face 56a
of the valve body 56 is divided into a radially inner portion and a
radially outer portion by an imaginary cylinder, which is shown by
broken lines in FIG. 4. The imaginary cylinder corresponds to the
wall defining the through hole 60. The pressure receiving area of
the radially inner portion is expressed by S1-S3, and that of the
radially outer portion is expressed by S4-S1. The radially inner
portion receives a downward force based on the pressure in the
through hole 60 (crank pressure Pc). The radially outer portion
receives a downward force based on the pressure PdL in the valve
chamber 59.
[0053] As described above, the pressure PdL in the valve chamber 59
is applied to the plunger chamber 71. The upper surface 72a of the
plunger 72 has a pressure receiving area that is equal to that of
the lower surface 72b (see FIG. 3), and the forces that act on the
plunger 72b ased on the pressure PdL offset each other. However,
the lower end face 57a of the guide 57 receives an upward force
based on the pressure PdL in the plunger chamber 71. The effective
pressure receiving area of the lower end face 57 a is equal to the
cross-sectional area S4 of the guide 57. Provided that the upward
forces are positive values, the net force .SIGMA.F2 acting upon the
guide 57 can be expressed by the following equation II. 1 F2 = F +
f2 - Pc ( S1 - S3 ) - PdL ( S4 - S1 ) + PdL S4 = F + f2 + PdL S1 -
Pc ( S1 - S3 ) Equation II
[0054] In the process of simplifying equation II, -PdL.multidot.S4
is canceled by +PdL.multidot.S4 , and the term +PdL.multidot.S1
remains. Thus, the resultant of the downward force based on the
pressure PdL acting upon the guide 57 and the upward force based on
the pressure PdL acting upon the guide 57 is a net upward force,
and the magnitude of this resultant force depends only on the
cross-sectional area S1 of the through hole 60. The surface area of
the portion of the guide 57 that receives the pressure PdL with
effect, i.e., the effective pressure receiving area of the guide 57
with respect to the pressure PdL, is always equal to the
cross-sectional area S1 of the through hole 60 regardless of the
cross-sectional area S4 of the guide 57 and the cross-sectional
area of the plunger 72.
[0055] The axial position of the operating rod 53 is determined
such that the force .SIGMA.F1 in the equation I and the force EF 2
in the equation II are equal. When the force .SIGMA.F1 is equal to
the force .SIGMA.F2 (.SIGMA.F1=.SIGMA.F2), the following equation
III is satisfied.
PdH.multidot.S2-PdL.multidot.S1-Pc (S2-S1)=F-f1+f2 Equation III
[0056] The cross-sectional area S1 of the through hole 60 is equal
to the cross-sectional area S2 of the divider 54. Therefore, if S2
is replaced with S1 in equation III, the following equation IV is
obtained.
PdH-PdL=(F-f1+f2)/S1 Equation IV
[0057] In equation IV, f1, f2 and S1 are determined by the design
of the control valve 100. The electromagnetic force F is a variable
parameter that changes depending on the power supplied to the coil
75. The equation IV shows that the operating rod 53 operates to
change the pressure difference (PdH-PdL) in accordance with the
change in the electromagnetic force F. In other words, the
operating rod 53 operates in accordance with the pressure PdH and
the pressure PdL, which act on the rod 53, such that the pressure
difference (PdH-PdL) seeks a target value, which is determined by
the electromagnetic force F. The operating rod 53 and the plunger
72 function as a pressure detecting body or a pressure receiving
body.
[0058] As described above, the downward force f1 of the return
spring 68 is greater than the upward force f2 of the shock
absorbing spring 74. Therefore, when no voltage is applied to the
coil 75, or when the electromagnetic force F is nil, the operating
rod 53 moves to the initial position shown in FIG. 3 to maximize
the opening size of the through hole 60 by the valve body 56.
[0059] When the duty ratio Dt of the voltage applied to the coil 75
is the minimum value Dt(min) in a preset range, the upward
electromagnetic force F exceeds the downward force f1 of the return
spring 68. The upward urging force F and the upward force f2 of the
shock absorbing spring 74 compete with the downward force f1 of the
return spring 68 and the downward force based on the pressure
difference (PdH-PdL). The operating rod 53 operates to satisfy the
above equation IV to determine the position of the valve body 56
with respect to the valve seat 64. Then, refrigerant gas is
supplied, from the second pressure monitoring point P2, through the
supply passage 110 to the crank chamber 15 at a flow rate that
depends on the valve position of the valve body 56, to adjust the
crank pressure Pc.
[0060] As shown in FIGS. 2 and 3, the controller 81 is a computer,
which includes a CPU, a ROM, a RAM and an input-output interface.
Detectors 83 detect various external information necessary for
controlling the compressor and send the information to the
controller 81. The controller 81 computes an appropriate duty ratio
Dt based on the information and commands the drive circuit 82 to
output a voltage having the computed duty ratio Dt. The drive
circuit 82 outputs the instructed pulse voltage having the duty
ratio Dt to the coil 75 of the control valve 100. The
electromagnetic force F of the solenoid 52 is determined according
to the duty ratio Dt.
[0061] The detectors 83 may include, for example, an air
conditioner switch, a passenger compartment temperature sensor, a
temperature adjuster for setting a desired temperature in the
passenger compartment, and a throttle sensor for detecting the
opening size of a throttle valve of the engine Eg. The detectors 83
may also include a pedal position sensor for detecting the
depression degree of an acceleration pedal of the vehicle. The
opening size of the throttle valve and the depression degree of the
acceleration pedal represent the load on the engine Eg.
[0062] The flowchart of FIG. 5 shows the main routine for
controlling the compressor displacement. When the vehicle ignition
switch or the starting switch is turned on, the controller 81
starts processing. The controller 81 performs various initial
setting in step S41. For example, the controller 81 assigns
predetermined initial value to the duty ratio Dt of the voltage
applied to the coil 75.
[0063] In step S42, the controller 81 waits until the air
conditioner switch is turned on. When the air conditioner switch is
turned on, the controller 81 moves to step S43. In step S43, the
controller 81 judges whether the vehicle is in an exceptional
driving mode. The exceptional driving mode refers to, for example,
a case where the engine Eg is under high-load conditions such as
when driving uphill or when accelerating rapidly. The controller 81
judges whether the vehicle is in the exceptional driving mode
according to, for example, external information from the throttle
sensor or the pedal position sensor.
[0064] If the outcome of step S43 is negative, the controller 81
judges that the vehicle is in a normal driving mode and moves to
step S44. The controller 81 then executes a normal control
procedure shown in FIG. 6. If the outcome of step S43 is positive,
the controller 81 executes an exceptional control procedure for
temporarily limiting the compressor displacement in step S45. The
exceptional control procedure differs according to the nature of
the exceptional driving mode. FIG. 7 illustrates an example of the
exceptional control procedure that is executed when the vehicle is
rapidly accelerated.
[0065] The normal control procedure of FIG. 6 will now be
described. In step S51, the controller 81 judges whether the
temperature Te(t), which is detected by the temperature sensor, is
higher than a desired temperature Te(set), which is set by the
temperature adjuster. If the outcome of step S51 is negative, the
controller 81 moves to step S52. In step S52, the controller 81
judges whether the temperature Te(t) is lower than the desired
temperature Te(set). If the outcome in step S52 is also negative,
the controller 81 judges that the detected temperature Te(t) is
equal to the desired temperature Te(set) and returns to the main
routine of FIG. 5 without changing the current duty ratio Dt.
[0066] If the outcome of step S51 is positive, the controller 81
moves to step S53 for increasing the cooling performance of the
refrigerant circuit. In step S53, the controller 81 adds a
predetermined value .DELTA.D to the current duty ratio Dt and sets
the resultant as a new duty ratio Dt. The controller 81 sends the
new duty ratio Dt to the drive circuit 82. Accordingly, the
electromagnetic force F of the solenoid 52 is increased by an
amount that corresponds to the value .DELTA.D, which moves the rod
53 in the valve closing direction. As the rod 53 moves, the force
f1 of the return spring 68 is increased. The axial position of the
rod 53 is determined such that equation IV is satisfied.
[0067] As a result, the opening size of the control valve 100 is
decreased and the crank pressure Pc is lowered. Thus, the
inclination angle of the swash plate 18 and the compressor
displacement are increased. An increase of the compressor
displacement increases the flow rate of refrigerant in the
refrigerant circuit and increases the cooling performance of the
evaporator 38. Accordingly, the temperature Te(t) is lowered to the
desired temperature Te(set) and the pressure difference (PdH-PdL)
is increased.
[0068] If the outcome of S52 is positive, the controller 81 moves
to step S54 for decreasing the cooling performance of the
refrigerant circuit. In step S54, the controller 81 subtracts the
predetermined value .DELTA.D from the current duty ratio Dt and
sets the resultant as a new duty ratio Dt. The controller 81 sends
the new duty ratio Dt to the drive circuit 82. Accordingly, the
electromagnetic force F of the solenoid 52 is decreased by an
amount that corresponds to the value .DELTA.D, which moves the rod
53 in the valve opening direction. As the rod 53 moves, the force
f1 of the return spring 68 is decreased. The axial position of the
rod 53 is determined such that equation IV is satisfied.
[0069] As a result, the opening size of the control valve 100 is
increased and the crank pressure Pc is raised. Thus, the
inclination angle of the swash plate 18 and the compressor
displacement are decreased. A decrease of the compressor
displacement decreases the flow rate of refrigerant in the
refrigerant circuit and decreases the cooling performance of the
evaporator 38. Accordingly, the temperature Te(t) is raised to the
desired temperature Te(set) and the pressure difference (PdH-PdL)
is decreased.
[0070] As described above, the duty ratio Dt is optimized in steps
S53 and S54 such that the detected temperature Te(t) seeks the
desired temperature Te(set).
[0071] The exceptional control procedure of FIG. 7 will now be
described. In step S81, the controller 81 stores the current duty
ratio Dt as a restoration target value DtR. In step S82, the
controller 81 stores the current detected temperature Te(t) as an
initial temperature Te(INI), or the temperature when the
displacement limiting control procedure is started.
[0072] In step S83, the controller 81 starts a timer. In step S84,
the controller 81 changes the duty ratio Dt to zero percent and
stops applying voltage to the coil 75. Accordingly, the opening
size of the control valve 100 is maximized by the return spring 68,
which increases the crank pressure Pc and minimizes the compressor
displacement. As a result, the torque of the compressor is
decreased, which reduces the load on the engine Eg when the vehicle
is rapidly accelerated.
[0073] In step S85, the controller 81 judges whether the elapsed
period STM measured by the timer is more than a predetermined
period ST. Until the measured period STM surpasses the
predetermined period ST, the controller 81 maintains the duty ratio
Dt at zero percent. Therefore, the compressor displacement and
torque are maintained at the minimum levels until the predetermined
period ST elapses. The predetermined period ST starts when the
displacement limiting control procedure is started. This permits
the vehicle to be smoothly accelerated. Since acceleration is
generally temporary, the period ST need not be long.
[0074] When the measured period STM surpasses the period ST, the
controller 81 moves to step S86. In step S86, the controller 81
judges whether the current temperature Te(t) is higher than a value
computed by adding a value .beta. to the initial temperature
Te(INI). If the outcome of step S86 is negative, the controller 81
judges that the compartment temperature is in an acceptable range
and maintains the duty ratio Dt at zero percent. If the outcome of
step S86 is positive, the controller 81 judges that the compartment
temperature has increased above the acceptable range due to the
displacement limiting control procedure. In this case, the
controller 81 moves to step S87 and restores the cooling
performance of the refrigerant circuit.
[0075] In step S87, the controller 81 executes a duty ratio
restoration control procedure. In this procedure, the duty ratio Dt
is gradually restored to the restoration target value DtR over a
certain period. Therefore, the inclination of the swash plate 18 is
changed gradually, which prevents the shock of a rapid change. In
the chart of step S87, the period from time t3 to time t4
represents a period from when the duty ratio Dt is set to zero
percent in step S84 to when the outcome of step S86 is judged to be
positive. The duty ratio Dt is restored to the restoration target
value DtR from zero percent over the period from the time t4 to
time t5. When the duty ratio Dt reaches the restoration target
value DtR, the controller 81 moves to the main routine shown in
FIG. 5.
[0076] This embodiment has the following advantages.
[0077] The control valve 100 does not directly control the suction
pressure Ps, which is influenced by the thermal load on the
evaporator 38. The control valve 100 directly controls the pressure
difference (PdH-PdL) between the pressures at the pressure
monitoring points P1, P2 in the refrigerant circuit for controlling
the compressor displacement. Therefore, the compressor displacement
is controlled regardless of the thermal load on the evaporator 38.
During the exceptional control procedure, no voltage is applied to
the control valve 100, which quickly minimizes the compressor
displacement. Accordingly, during the exceptional control
procedure, the displacement is limited and the engine load is
decreased. The vehicle therefore runs smoothly.
[0078] During the normal control procedure, the duty ratio Dt is
adjusted based on the detected temperature Te(t) and the desired
temperature Te(set), and the operating rod 53 operates depending on
the pressure difference (PdH-PdL). That is, the control valve 100
not only operates based on external commands but also automatically
operates in accordance with the pressure difference (PdH-PdL),
which acts on the control valve 100. The control valve 100
therefore effectively controls the compressor displacement such
that the actual temperature Te(t) seeks the target temperature
Te(set) and maintains the target temperature Te(set) in a stable
manner. Further, the control valve 200 quickly changes the
compressor displacement when necessary.
[0079] The duty ratio Dt of the voltage applied to the solenoid 52,
i.e., the electromagnetic force F of the solenoid 52, indicates the
desired value of the pressure difference (PdH-PdL). The operating
rod 53 operates according to the pressure difference (PdH-PdL) so
that the pressure difference (PdH-PdL) is steered to the desired
value. Thus, the intended displacement control is constantly and
reliably realized. For example, when the compressor is operating at
the minimum displacement in the exceptional control procedure, the
compressor can easily return to a normal displacement according to
a desired recovery pattern, and such a recovery pattern is easily
set to avoid shocks that may occur due to the displacement
increase.
[0080] The second pressure introduction passage 42 for connecting
the second pressure monitoring point P2 to the control valve 100
functions as a part of the supply passage 110. Therefore, the
second pressure introduction passage 42 need not be formed
separately from the supply passage 110. This simplifies the
compressor and the control valve 110. That is, the number of
passages formed in the compressor is minimized. Also, the number of
ports formed in the control valve 100 and the number of seal rings
used in the control valve 100 are minimized.
[0081] The operating rod 53 integrally includes the divider 54, the
coupler 55 and the guide 57 in a single body, and a part of the
guide 57 forms the valve body 56. This reduces the number of parts
and simplifies the control valve 100.
[0082] The pressure acting on the operating rod 53 includes the
pressure PdH at the first pressure monitoring point P1, 320 the
pressure PdL at the second pressure monitoring point and the crank
pressure Pc. However, as can be understood from the above equation
IV, the force based on the crank pressure Pc has substantially no
effect on the operating rod 53. This is mainly because the
cross-sectional area S1 of the through hole 60, more specifically,
the cross-sectional area S1 of the portion of the through hole 60
opening to the valve chamber 59, is the same as the cross-sectional
area S2 of the divider 54. Therefore, the gas pressures determining
the axial position of the operating rod 53 are only the pressure
PdH at the first pressure monitoring point P1 and the pressure PdL
at the second pressure monitoring point P2. This allows the
operating rod 53 to operate smoothly depending on the pressure
difference (PdH-PdL) under no and not the crank pressure Pc, thus
producing a highly accurate displacement control valve.
[0083] The diameter of the through hole 60 is constant in the axial
direction and is equal to the diameter of the divider 54. Thus, in
assembling the control valve 100, the operating rod 53 as an
integral body and can be inserted easily into the through hole 60
from the valve chamber 59 side.
[0084] It should be apparent to those skilled in the art that the
present invention may be embodied in many other specific forms
without departing from the spirit or scope of the invention.
Particularly, it should be understood that the invention may be
embodied in the following forms.
[0085] FIG. 8 shows a control valve 100 according to a second
embodiment of the present invention. FIG. 9 shows a control valve
100 according to a third embodiment of the present invention. In
each of these control valves 100, the supply passage 110 is defined
by the first pressure introduction passage 41 and the crank passage
44. Accordingly, the internal constructions of the control valves
100 are changed somewhat, as shown in FIGS. 8 and 9, respectively,
compared with the control valve 100 shown in FIG. 3. The same or
like components have the same reference numbers in all
embodiments.
[0086] Since the control valve of FIG. 8 is basically the same as
that of the control valve 100 of FIG. 3, further description of it
will be omitted.
[0087] In the control valve 100 of FIG. 9, a clearance (not shown)
is defined between the plunger 72 and the receiving cylinder 69.
This clearance permits application of the pressure PdH to the
plunger chamber 71.
[0088] Further, in the control valve of FIG. 9, the positional
relationship between the plunger 72 and the fixed iron core 70 is
reversed compared with the control valves 100 in FIGS. 3 and 8. The
valve body 56 is not integrated with the operating rod 3 but is
independent. However, the electromagnetic force of the solenoid 52
acts against the operating rod 53 in the valve closing direction
like in the control valves shown in FIGS. 3 and 8.
[0089] Unlike the control valves 100 shown in FIGS. 3 and 8, the
force of the return spring 68 is weaker than the force of the shock
absorbing spring 74. When no voltage is applied to the coil 75, the
shock absorbing spring 74 moves the plunger 72 and the operating
rod 53 in the valve opening direction. Thus, the valve body 56
opens the through hole 60 fully, as shown in FIG. 9. The
electromagnetic force generated between the plunger 72 and the
fixed iron core 70, when a voltage is applied to the coil 75, moves
the operating rod 53 in the valve closing direction. Since the
return spring 68 presses the valve body 56 against the operating
rod 53, the valve body 56 moves integrally with the operating rod
53.
[0090] In a fourth embodiment shown in FIG. 10, the aligning
mechanism including the spring seat 79 and the aligning ball 80 of
the control valve 100 shown in FIG. 3 is omitted. The return spring
68 is directly abutted against the divider 54 of the operating rod
53. The divider 54 has at the upper end a boss 54a for receiving
the return spring 68.
[0091] In the control valve 100 of FIG. 3, the cross-sectional area
S1 of the portion of the through hole 60 opening to the valve
chamber 59 may be smaller than the cross-sectional area S2 of the
divider 54. The merits of such a control valve 100 will be
described. The following equation V is a modification of the above
equation III. In equation V, S1 is smaller than S2.
(PdH-Pc)S2-(PdL-Pc)S1=F-f1+f2 Equation V
[0092] When equation IV is rearranged so that the right side in
equation IV is equal to that of equation V, the following equation
VI is obtained.
(PdH-PdL)S1=F-f1+f2 Equation VI
[0093] When the left side in equation V is compared with that in
equation VI, under the condition of PdH>PdL>Pc, the following
relationship is established.
(PdH-Pc)S2-(PdL-Pc)S1>(PdH-PdL)S1
[0094] Thus, when the control valve 100 satisfies the condition
S2>S1, the force based on the pressure difference (PdH-PdL) that
acts on the operating rod 53 is greater than that when S2=S1.
Therefore, when S2>S1, even if the flow rate of the refrigerant
in the refrigerant circuit is relatively low, i.e., even if the
pressure difference (PdH-PdL) is relatively small, the pressure
difference (PdH-PdL) reliably determines the position of the
operating rod 53.
[0095] The control valve 100 may be designed to adjust the aperture
size of the bleed passage 31 in addition to that of the supply
passage 110.
[0096] The first pressure monitoring point P1 need not be located
in the discharge chamber 23. The first pressure monitoring point P1
may be located at any position as long as the position is exposed
to the discharge pressure Pd. In other words, the first pressure
monitoring point P1 may be located anywhere in a high pressure zone
of the refrigerant circuit, which includes the discharge chamber
23, the condenser 36 and the higher pressure pipe 40. The second
pressure monitoring point P2 may be located at any position that is
downstream of the first pressure monitoring point P1 in the high
pressure zone.
[0097] The present invention can be embodied in a control valve of
a wobble type variable displacement compressor.
[0098] Therefore, the present examples and embodiments are to be
considered as illustrative and not restrictive and the invention is
not to be limited to the details given herein, but may be modified
within the scope and equivalence of the appended claims.
* * * * *