U.S. patent application number 09/745092 was filed with the patent office on 2001-08-16 for displacement control apparatus and method for variable displacement compressor.
Invention is credited to Adaniya, Taku, Kawaguchi, Masahiro, Kimura, Kazuya, Matsubara, Ryo, Ota, Masaki, Suitou, Ken.
Application Number | 20010013225 09/745092 |
Document ID | / |
Family ID | 18490754 |
Filed Date | 2001-08-16 |
United States Patent
Application |
20010013225 |
Kind Code |
A1 |
Ota, Masaki ; et
al. |
August 16, 2001 |
Displacement control apparatus and method for variable displacement
compressor
Abstract
A control valve is located in a variable displacement compressor
used in a refrigerant circuit. The control valve operates such that
the pressure difference between two pressure monitoring points in
the refrigerant circuit seeks a predetermined target value. A
controller determines the target value of the pressure difference
in accordance with external information that represents the
required cooling performance. The target value of the pressure
difference is represented by a duty ratio applied to the control
valve. When the acceleration pedal is pressed beyond a
predetermined level, the controller limits the duty ratio.
Therefore, the compressor torque does not hinder quick vehicle
acceleration. Also, the cooling performance is not lowered unless
it is necessary.
Inventors: |
Ota, Masaki; (Kariya-shi,
Aichi-ken, JP) ; Kimura, Kazuya; (Kariya-shi,
Aichi-ken, JP) ; Kawaguchi, Masahiro; (Kariya-shi,
Aichi-ken, JP) ; Suitou, Ken; (Kariya-shi, Aichi-ken,
JP) ; Matsubara, Ryo; (Kariya-shi, Aichi-ken, JP)
; Adaniya, Taku; (Kariya-shi, Aichi-ken, JP) |
Correspondence
Address: |
MORGAN & FINNEGAN, L.L.P.
345 Park Avenue
New York
NY
10154
US
|
Family ID: |
18490754 |
Appl. No.: |
09/745092 |
Filed: |
December 20, 2000 |
Current U.S.
Class: |
62/228.5 ;
417/222.2 |
Current CPC
Class: |
F04B 27/1804 20130101;
F04B 2027/1854 20130101; F04B 2027/1813 20130101; F04B 2205/07
20130101 |
Class at
Publication: |
62/228.5 ;
417/222.2 |
International
Class: |
F25B 001/00 |
Foreign Application Data
Date |
Code |
Application Number |
Dec 24, 1999 |
JP |
11-368009 |
Claims
What is claimed is:
1. A displacement control apparatus for a variable displacement
compressor used in a refrigerant circuit of a vehicle air
conditioner, wherein the compressor is driven by a drive source of
a vehicle, the apparatus comprising: a displacement control
mechanism, which controls the displacement of the compressor based
on the pressure difference between the pressures at two pressure
monitoring points located in the refrigerant circuit, the pressure
difference representing the displacement of the compressor; a first
device for detecting external information representing the required
cooling performance of the refrigerant circuit; a second device for
detecting external information representing the load acting on the
drive source; and a controller for determining a target value of
the pressure difference based on the external information detected
by the first device, wherein the displacement control mechanism
controls the displacement of the compressor such that the pressure
difference seeks the target value, wherein the controller judges
whether to set a limit value of the pressure difference based on
the external information detected by the second device, wherein,
when the limit value is set and a compressor displacement that
corresponds to the target value is greater than a compressor
displacement that corresponds to the limit value, the controller
uses the limit value as the target value of the pressure difference
to limit the compressor displacement.
2. The displacement control apparatus according to claim 1, wherein
the controller sets the limit value when the load on the drive
source is equal to or greater than a predetermined level.
3. The displacement control apparatus according to claim 1, wherein
the controller changes the limit value in accordance with the load
acting on the drive source.
4. The displacement control apparatus according to claim 3, wherein
the controller discretely changes the limit value in accordance
with the load acting on the drive source.
5. The displacement control apparatus according to claim 1, wherein
the controller maintains the target value of the pressure
difference at the limit value for a predetermined period and then
changes the target value to a target value that is determined based
on the external information detected by the first device over a
predetermined period.
6. The displacement control apparatus according to claim 1, wherein
the compressor includes a crank chamber, an inclining drive plate
located in the crank chamber and a piston, which is reciprocated by
the drive plate, wherein the inclination angle of the drive plate
changes in accordance with the pressure in the crank chamber, and
the inclination angle of the drive plate determines the stroke of
the piston and the compressor displacement, wherein the
displacement control mechanism includes a control valve located in
the compressor, and wherein the control valve operates depending on
the pressure difference to adjust the pressure in the crank
chamber.
7. The displacement control apparatus according to claim 6, wherein
the control valve includes: a valve body; an actuator for urging
the valve body, wherein the controller controls power supplied to
the actuator such that the urging force of the actuator corresponds
to the target value; and a pressure receiving body, wherein the
pressure receiving body actuates the valve body in accordance with
the pressure difference acting on the pressure receiving body such
that the pressure difference seeks the target value.
8. The displacement control apparatus according to claim 1, wherein
the first device detects external information related to a
temperature.
9. The displacement control apparatus according to claim 8, wherein
the first device includes a temperature sensor for detecting the
temperature in the passenger compartment and a temperature adjuster
for setting a target value of the compartment temperature, and
wherein the controller determines the target value of the pressure
difference based on the difference between the detected compartment
temperature and the target temperature.
10. The displacement control apparatus according to claim 1,
wherein the second device includes a pedal position sensor for
detecting the depression degree of an acceleration pedal of the
vehicle.
11. A displacement control apparatus for a variable displacement
compressor used in a refrigerant circuit of a vehicle air
conditioner, wherein the compressor is driven by a drive source of
a vehicle, the apparatus comprising: a displacement control
mechanism, which controls the displacement of the compressor based
on the pressure difference between the pressures at two pressure
monitoring points located in the refrigerant circuit, the pressure
difference representing the displacement of the compressor; a first
device for detecting external information representing the required
cooling performance of the refrigerant circuit; a second device for
detecting information representing the load acting on the external
drive source; a determining means for determining a target value of
the pressure difference in accordance with the external information
detected by the first device, wherein the displacement control
mechanism controls the displacement of the compressor such that the
pressure difference seeks the target value; a judging means for
judging whether to set a limit value of the pressure difference
based on the external information detected by the second device,
and a means for using the limit value as the target value of the
pressure difference to limit the compressor displacement when the
limit value is set and a compressor displacement that corresponds
to the target value is greater than a compressor displacement that
corresponds to the limit value.
12. A method for controlling the displacement of a variable
displacement compressor used in a refrigerant circuit of a vehicle
air conditioner, wherein the compressor is driven by a drive source
of a vehicle, the method comprising: determining a target value of
the pressure difference between the pressures at two pressure
monitoring points located in the refrigerant circuit based on
external information that represents the required cooling
performance of the refrigerant circuit, the pressure difference
representing the displacement of the compressor; controlling the
compressor displacement such that the pressure difference seeks the
target value; judging whether to set a limit value of the pressure
difference based on external information that represents the load
acting on the drive source; and using the limit value as the target
value of the pressure difference when the limit value is set and
when a compressor displacement that corresponds to the target value
is greater than a compressor displacement that corresponds to the
limit value.
13. The method according to claim 12, wherein, when the load acting
on the drive source is equal to or greater than a predetermined
level, the limit value is set.
14. The method according to claim 12, further comprising: changing
the limit value in accordance with the load acting on the drive
source.
15. The method according to claim 14, wherein the limit value is
discretely changed in accordance with the load acting on the drive
source.
16. The method according to claim 12, further comprising:
maintaining the target value of the pressure difference at the
limit value for a predetermined period; and changing the target
value to a target value that is determined based on the external
information representing the required cooling performance of the
refrigerant circuit over a predetermined period.
Description
BACKGROUND OF THE INVENTION
[0001] The present invention relates to a variable displacement
compressor used in a refrigerant circuit of a vehicle air
conditioner. More particularly, the present invention pertains to a
displacement control apparatus and a displacement control method
for the variable displacement compressor.
[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. 8 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. 8) 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 displacement control apparatus and a displacement control
method for 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 displacement control apparatus for a variable
displacement compressor used in a refrigerant circuit of a vehicle
air conditioner. The compressor is driven by a drive source of a
vehicle. The apparatus includes a displacement control apparatus, a
first device, a second device and a controller. The displacement
control mechanism controls the displacement of the compressor based
on the pressure difference between the pressures at two pressure
monitoring points located in the refrigerant circuit. The pressure
difference represents the displacement of the compressor. The first
device detects external information representing the required
cooling performance of the refrigerant circuit. The second device
detects external information representing the load acting on the
drive source. The controller determines a target value of the
pressure difference based on the external information detected by
the first device. The displacement control mechanism controls the
displacement of the compressor such that the pressure difference
seeks the target value and judges whether to set a limit value of
the pressure difference based on the external information detected
by the second device. When the limit value is set and a compressor
displacement that corresponds to the target value is greater than a
compressor displacement that corresponds to the limit value, the
controller uses the limit value as the target value of the pressure
difference to limit the compressor displacement.
[0010] The present invention may also be embodied in a method for
controlling the displacement of a variable displacement compressor
used in a refrigerant circuit of a vehicle air conditioner. The
compressor is driven by a drive source of a vehicle. The method
includes determining a target value of the pressure difference
between the pressures at two pressure monitoring points located in
the refrigerant circuit based on external information that
represents the required cooling performance of the refrigerant
circuit, the pressure difference representing the displacement of
the compressor, controlling the compressor displacement such that
the pressure difference seeks the target value, judging whether to
set a limit value of the pressure difference based on external
information that represents the load acting on the drive source,
and using the limit value as the target value of the pressure
difference when the limit value is set and when a compressor
displacement that corresponds to the target value is greater than a
compressor displacement that corresponds to the limit value.
[0011] 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
[0012] 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:
[0013] FIG. 1 is a cross-sectional view illustrating a variable
displacement swash plate type compressor according to one
embodiment of the present invention;
[0014] FIG. 2 is a schematic diagram illustrating a refrigerant
circuit including the compressor of FIG. 1;
[0015] FIG. 3 is a cross-sectional view illustrating a control
valve of FIG. 1;
[0016] FIG. 4 is a schematic cross-sectional view showing part of
the control valve shown in FIG. 3;
[0017] FIG. 5 is a flowchart showing a main routine for controlling
a compressor displacement;
[0018] FIG. 6 is a flowchart showing a normal control
procedure;
[0019] FIG. 7 is a flow chart showing an exceptional control
procedure; and
[0020] FIG. 8 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
[0021] One 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.
[0022] 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.
[0023] The front end of the drive shaft 16 is connected to an
external drive source, which is an internal combustion engine Eg
used in a vehicle 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.
[0024] 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.
[0025] 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.
[0026] 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.
[0027] 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.
[0028] 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.
[0029] 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.
The speed of the drive shaft 16 is computed based on the speed of
the engine Eg and the ratio of the speed of the drive shaft 16 to
the speed of the engine Eg. The speed ratio is determined by the
power transmission mechanism PT. 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.
[0030] 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 PI, 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.
[0031] 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 46.
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 PI, to the control valve 46. The second pressure
introduction passage 42 connects the second pressure monitoring
point P2 to the control valve 46. The crank passage 44 connects the
control valve 46 to the crank chamber 15.
[0032] 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 46 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.
[0033] 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 PI 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 46 in the compressor, can
be shortened.
[0034] As shown in FIG. 1, the control valve 46 is fitted in a
receiving hole 14a of the rear housing member 14. As shown in FIGS.
3 and 4, the control valve 46 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.
[0035] The control valve 46 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.
[0036] A first radial port 62 is formed in the upper housing member
58b to communicate with the valve chamber 59. 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 46.
[0037] 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.
[0038] 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.
[0039] 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.
[0040] 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.
[0041] 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 the 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.
[0042] 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.
[0043] 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.
[0044] 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.
[0045] 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.
[0046] 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).
[0047] 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.EF acting upon the divider 54 can be expressed
by the following equation I.
.SIGMA.F1=PdH.multidot.S2-Pc(S2-S3)+f1 Equation I
[0048] 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.
[0049] 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 72 based 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 57a 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.
.SIGMA.F2=F+f2-Pc(S1-S3)-PdL(S4-S1)+PdL.multidot.S4=F+f2+PdL.multidot.S1-P-
c(S1-S3) Equation II
[0050] 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.
[0051] The axial position of the operating rod 53 is determined
such that the force .SIGMA.F1 in the equation I and the force
.SIGMA.F2 in the equation II are equal. When the force EF1 is equal
to the force .SIGMA.ZF2 (.SIGMA.F1=.SIGMA.F2), the following
equation III is satisfied.
PdH.multidot.S2-PdL.multidot.S1-Pc(S2-S1)=F-f1+f2 Equation III
[0052] 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
[0053] In equation IV, f1, f2 and S1 are determined by the design
of the control valve 46. 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 functions as a
pressure detecting body or a pressure receiving body.
[0054] 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.
Several devices 83 to 86 detect various external information
necessary for controlling the compressor and send the information
to the controller 81. The devices 83 to 86 include an air
conditioner switch 83, a passenger compartment temperature sensor
84, a temperature adjuster 85 for setting a desired temperature in
the passenger compartment and a pedal position sensor 86 for
detecting the depression degree of an acceleration pedal of the
vehicle. Instead of or in addition to the pedal position sensor 86,
the devices may include a throttle sensor for detecting the opening
size of a throttle valve of the engine Eg. The temperature sensor
84 and the temperature adjuster 85 detect external information
representing the required cooling performance of the refrigerant
circuit. The depression degree of the acceleration pedal and the
opening size of the throttle valve represent the load on the engine
Eg.
[0055] The controller 81 computes an appropriate duty ratio Dt
based on the information from the devices 83 to 86 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 46.
The electromagnetic force F of the solenoid 52 is determined
according to the duty ratio Dt.
[0056] 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
settings in step S101. For example, the controller 81 assigns
predetermined initial value to the duty ratio Dt of the voltage
applied to the coil 75.
[0057] In step S102, the controller 81 waits until the air
conditioner switch 83 is turned on. When the air conditioner switch
83 is turned on, the controller sets a limit value Dtlm of the duty
ratio Dt in accordance with the load on the engine Eg in steps S103
to S108.
[0058] In step S103, the controller 81 judges whether the pedal
depression degree detected by the pedal position sensor 86 is equal
to or greater than a predetermined first value ACC1. If the outcome
of step S103 is negative, the controller 81 judges that the pedal
depression degree ACC is relatively small, or that there is no
demand for quick acceleration. In this case, the compressor
displacement need not be limited for decreasing the load on the
engine Eg. Therefore, the controller 81 does not set the limit
value Dtlm of the duty ratio Dt and moves to step S110.
[0059] If the outcome of step S103 is positive, the controller 81
judges that there is a demand for quick acceleration and sets the
limit value Dtlm of the duty ratio Dt. The limit value Dtlm is set
in accordance with the pedal depression degree ACC, or the degree
of the required acceleration (engine load).
[0060] In step S104, the controller 81 judges whether the pedal
depression degree ACC is equal to or greater than a predetermined
second value ACC2. The second value ACC2 is greater than the first
value ACC1. If the outcome of step S104 is negative, that is, if an
inequality ACC1.ltoreq.ACC<ACC2 is satisfied, the controller 81
judges that the required degree of acceleration or the engine load
is relatively small and moves to step S105. In step S105, the
controller 81 sets the limit value Dtlm to a predetermined first
value Dtlm1 moves to step S109.
[0061] If the outcome of step S104 is positive, the controller 81
moves to step S106 and judges whether the pedal depression degree
ACC is equal to or greater than a predetermined third value ACC3.
The third value ACC is greater than the second value ACC2. If the
outcome of step S106 is negative, that is, if an inequality
ACC2.ltoreq.ACC<ACC3 is satisfied, the controller 81 judges that
the degree of the required quick acceleration is intermediate and
moves to step S107. In step S107, the controller 81 sets the limit
value Dtlm to a predetermined second value Dtlm2 and moves to step
S109. The second value Dtlm2 is less than the first value
Dtlm1.
[0062] If the outcome of step S106 is positive, the controller 81
judges that the degree of the required quick acceleration or the
engine load is relatively great and moves to step S108. In step
S108, the controller 81 sets the limit value Dtlm to a
predetermined third value Dtlm3 and moves to step S109. The third
value Dtlm3 is less than the second value Dtlm2 and is, for
example, zero percent.
[0063] In step S109, the controller 81 judges whether the current
duty ratio Dt is greater than the limit value Dtlm, which is set in
accordance with the pedal depression degree ACC. In other words,
the controller 81 judges whether the compressor displacement that
corresponds to the current duty ratio Dt is greater than the
compressor displacement that corresponds to the limit value Dtlm.
The compressor displacement correlates with the compressor torque.
If the outcome of step S109 is negative, the controller 81 judges
that the compressor torque will not significantly increase the load
on the engine Eg during the currently required quick acceleration.
In step S110, the controller 81 executes a normal control procedure
shown in FIG. 6.
[0064] If the outcome of step S109 is positive, the controller 81
judges that the compressor torque will increase the load on the
engine Eg during the currently required quick acceleration and
moves to step S111. In step S111, the controller 81 executes an
exceptional control procedure shown in FIG. 7 for temporarily
limiting the compressor displacement and the compressor torque.
[0065] The normal control procedure of FIG. 6 will now be
described. In step S121, the controller 81 judges whether the
temperature Te(t), which is detected by the temperature sensor 84,
is higher than a desired temperature Te(set), which is set by the
temperature adjuster 85. If the outcome of step S121 is negative,
the controller 81 moves to step S122. In step S122, the controller
81 judges whether the temperature Te(t) is lower than the desired
temperature Te(set). If the outcome in step S122 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 S121 is positive, the controller 81
moves to step S123 for increasing the cooling performance of the
refrigerant circuit. In step S123, 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 46 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 S122 is positive, the controller 81 moves
to step S124 for decreasing the cooling performance of the
refrigerant circuit. In step S124, 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 46 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
S123 and S124 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 S131, the controller 81 stores the current duty
ratio Dt as a restoration target value DtR. In step S132, the
controller 81 starts a timer.
[0072] In step S133, the controller 81 sets the duty ratio Dt to
the limit value Dtlm, which was set in one of steps S105, S107 and
S108 of the main routine shown in FIG. 5. Therefore, the duty ratio
Dt is decreased to the limit value Dtlm. Accordingly, the
electromagnetic force of the solenoid 52 is decreased, which
increases the opening size of the control valve 46. As a result,
the inclination angle of the swash plate 18 and the compressor
displacement are decreased, which decreases the torque of the
compressor and reduces the engine load.
[0073] In step S134, 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 the limit value Dtlm. Therefore, the compressor displacement
and torque are limited 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 S135. In step S135, 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 S135,
the period from time t3 to time t4 represents a period from when
the duty ratio Dt is set to the limit value Dtlm in step S131 to
when the outcome of step S134 is judged to be positive. The duty
ratio Dt is restored to the restoration target value DtR from the
limit value Dtlm 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.
[0075] This embodiment has the following advantages.
[0076] The control valve 46 does not directly control the suction
pressure Ps, which is influenced by the thermal load on the
evaporator 38. The control valve 46 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, the voltage applied to
the control valve 46 is limited, which quickly limits the
compressor displacement. Accordingly, during the exceptional
control procedure, the displacement is limited and the engine load
is decreased. The vehicle therefore runs smoothly.
[0077] 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 46
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 46. The control valve 46 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 46 quickly changes the compressor
displacement when necessary.
[0078] 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 limited 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.
[0079] The pressures PdH and PdL at the pressure monitoring point
P1 and P2 need not be detected by electric sensors, which
simplifies the structure.
[0080] When the pedal depression degree ACC is less than the first
value ACC1, there is no demand for quick acceleration. In this
case, the limit value Dtlm of the duty ratio Dt is not set.
Therefore, the refrigerant circuit exerts its full cooling
performance for maintaining the compartment temperature at the
desired level.
[0081] When the pedal depression degree ACC is more than the first
value ACC1, there is a demand for quick acceleration. In this case,
the duty ratio Dt is set to the limit value Dtlm. The exceptional
control procedure is executed for limiting the compressor
displacement and the compressor torque only when the duty ratio Dt,
which is set in accordance with the required cooling performance of
the refrigerant circuit, exceeds the limit value Dtlm. In other
words, when there is a demand for quick acceleration, the
exceptional control procedure is not executed if the current
compressor torque does not significantly increase the load on the
engine Eg. The compressor displacement is limited only when the
compressor torque is judged to hinder quick acceleration. Thus, the
cooling performance is less frequently decreased to a level that is
lower than a required level.
[0082] The limit value Dtlm of the duty ratio Dt is set in
accordance with the pedal depression degree ACC, or the degree of
the need for quick acceleration (engine load). The smaller the
degree of the need for quick acceleration, the greater the limit
value Dtlm. Thus, compared to a case where the limit value Dtlm is
constant, whether the exceptional control procedure needs to be
executed is properly determined. Also, the exceptional control
procedure is less frequently executed. Therefore, the occasions in
which the cooling performance is lowered below a demanded level are
minimized.
[0083] During the exceptional control procedure, the duty ratio Dt
is set to the limit value Dtlm, which corresponds to the degree of
the need for quick acceleration. If the degree of the need for
quick acceleration is relatively small during the exceptional
control procedure, the duty ratio Dt is not greatly decreased,
which prevents the compressor displacement from being greatly
decreased. Therefore, compared to a case where the compressor
displacement is always minimized when the exceptional control
procedure is executed, the compressor displacement is not limited
more than required. During the exceptional control procedure, the
cooling performance is not lowered by an excessive degree.
[0084] Accordingly, the vehicle is quickly accelerated without
reducing the cooling performance by an excessive degree.
[0085] 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.
[0086] The limit value Dtlm need not be varied in accordance with
the pedal depression degree ACC. Instead, the limit value Dtlm may
be constant.
[0087] In the routine of FIG. 5, the degree of the demanded quick
acceleration, or the engine load, is judged to be in one of the
three regions. However, the engine load may be determined to be in
one of two regions or more than three regions.
[0088] In the duty ratio restoration control procedure shown in
FIG. 7, the duty ratio Dt may be discretely increased to the
restoration target value DtR.
[0089] In the duty ratio restoration control procedure shown in
FIG. 7, the duty ratio Dt may be quickly increased from the limit
value Dtlm to the restoration target value DtR if the difference
between the limit value Dtlm and the restoration target value DtR
is less than a predetermined value.
[0090] Instead of or in addition to the pedal depression degree
ACC, at least one of the following parameters, which represent the
engine load, may be used for judging the engine load. The
parameters include, for example, the changing speed of the pedal
depression degree ACC, the opening size of the throttle valve, the
flow rate of air that is drawn into the engine Eg, the pressure of
the air drawn into the engine Eq, the engine speed and the vehicle
speed.
[0091] For example, the engine load may be judged based on the
pedal depression degree ACC and the vehicle speed. In this case,
the engine load is judged to be relatively great due to an uphill
movement of the vehicle if the vehicle speed is low despite of a
relatively great value of the pedal depression degree ACC. In this
case, the routine similar to the routine of FIG. 5 is executed.
[0092] The devices for detecting external information that
represents the required level of the cooling performance may
include a solar radiation sensor and an external temperature sensor
instead of or in addition to the temperature sensor 84 and the
temperature adjuster 85.
[0093] 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 high 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.
[0094] The first pressure monitoring point P1 may be located in a
zone that is exposed to the suction pressure Ps (low pressure
zone), and the second pressure monitoring point P2 may be located
in a section of the low pressure zone that is downstream of the
first pressure monitoring point P1. The low pressure zone refers to
a section of the refrigerant circuit that includes the evaporator
38, the suction chamber 22 and the low pressure pipe 39.
[0095] The first pressure monitoring point P1 may be located in the
high pressure zone, and the second pressure monitoring point P2 may
be located in the low pressure zone.
[0096] The first pressure monitoring point P1 may be located in the
high pressure zone, and the second pressure monitoring point P2 may
be located in the crank chamber 15. Alternatively, the first
pressure monitoring point P1 may be located in the crank chamber
15, and the second pressure monitoring point P2 may be located in
the low pressure zone. The crank chamber 15 is an intermediate
pressure zone, which is exposed to a pressure that is lower than
the pressure of the high pressure zone and is higher than the
pressure of the low pressure zone.
[0097] The pressure monitoring points P1, P2 may be at any two
locations in the refrigerant circuit, which includes the compressor
and the external circuit 35.
[0098] If the first pressure monitoring point P1 is located in the
crank chamber 15 and the second pressure monitoring point P2 is
located in the low pressure zone, the pressure difference (Pc-Ps)
between the pressure monitoring points P1, P2 decreases as the
compressor displacement is increased, unlike the embodiment of
FIGS. 1 to 7. Thus, the limit value Dtlm of the duty ratio Dt is
set as a lower limit value, not an upper limit value. When the duty
ratio Dt, which represents the pressure difference (Pc-Ps), falls
below the limit value Dtlm, the duty ratio Dt is increased to the
limit value Dtlm.
[0099] The pressures PdH, PdL at the pressure monitoring points P1,
P2 may be detected by electric pressure sensors, respectively, and
the control valve 46 may be actuated in accordance with the
difference between the detected pressures.
[0100] The control valve 46 may be located in the bleed passage 31
to regulate the flow rate of gas released from the crank chamber 15
to the suction chamber 22.
[0101] The control valve 46 may be designed to adjust the aperture
size of the bleed passage 31 in addition to that of the supply
passage 110.
[0102] The power transmission mechanism PT may include a clutch
mechanism. In this case, if the pedal depression degree ACC is
greater than the third determination value ACC3, the clutch
mechanism may disconnect the compressor from the engine Eg.
[0103] The present invention can be embodied in a control valve of
a wobble type variable displacement compressor.
[0104] The drive source of the vehicle need not be an internal
combustion engine. The drive source may be an electric motor or a
hybrid engine, which includes an electric motor and an internal
combustion engine.
[0105] 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.
* * * * *