U.S. patent application number 09/755395 was filed with the patent office on 2001-07-12 for control device of variable displacement compressor.
Invention is credited to Kawaguchi, Masahiro, Kimura, Kazuya, Matsubara, Ryo, Suitou, Ken.
Application Number | 20010007194 09/755395 |
Document ID | / |
Family ID | 18530813 |
Filed Date | 2001-07-12 |
United States Patent
Application |
20010007194 |
Kind Code |
A1 |
Kimura, Kazuya ; et
al. |
July 12, 2001 |
Control device of variable displacement compressor
Abstract
A variable displacement compressor air conditions a compartment
and includes a suction chamber, a discharge chamber and a crank
chamber. A controller controls the pressure in the crank chamber to
vary the compressor displacement. Two pressure monitoring points
are located in a refrigerant circuit. The pressure in the crank
chamber is controlled by a control valve. The control valve
operates based on the pressure difference between the monitoring
points such that a target pressure difference is maintained. A
temperature sensor monitors the temperature of the compartment. A
detection circuit compares the monitored temperature with reference
values. When the monitored temperature surpasses one reference
value or falls below another, the detection circuit outputs a
detection signal. When receiving the detection signal, a pressure
difference changer gradually increases or decreases the target
value of the pressure differences accordingly.
Inventors: |
Kimura, Kazuya; (Kariya-shi,
JP) ; Kawaguchi, Masahiro; (Kariya-shi, JP) ;
Suitou, Ken; (Kariya-shi, JP) ; Matsubara, Ryo;
(Kariya-shi, JP) |
Correspondence
Address: |
Kurt E. Richter
Morgan & Finnegan, L.L.P.
345 Park Avenue
New York
NY
10154
US
|
Family ID: |
18530813 |
Appl. No.: |
09/755395 |
Filed: |
January 5, 2001 |
Current U.S.
Class: |
62/228.3 ;
62/228.1 |
Current CPC
Class: |
F04B 49/065 20130101;
F04B 2027/185 20130101; F04B 2205/07 20130101; F04B 2027/1813
20130101; F04B 27/1804 20130101; F04B 2207/03 20130101; F04B
2027/1827 20130101; F04B 2027/1854 20130101 |
Class at
Publication: |
62/228.3 ;
62/228.1 |
International
Class: |
F25B 001/00; F25B
049/00 |
Foreign Application Data
Date |
Code |
Application Number |
Jan 7, 2000 |
JP |
2000-001601 |
Claims
What is claimed is:
1. A controller for a variable displacement compressor, wherein the
compressor is used for air conditioning a compartment and includes
a suction pressure zone, a discharge pressure zone, and a control
chamber, which is connected to the suction pressure zone and to the
discharge pressure zone, and wherein the pressure in the control
chamber is adjusted for controlling the displacement of the
compressor, the controller comprising: a refrigerant circuit
connected to the compressor, wherein two pressure monitoring points
are located in the refrigerant circuit; a control valve for
controlling the pressure in the control chamber, wherein the
control valve operates based on the actual pressure difference
between the pressure monitoring points such that a target value of
the pressure difference between the pressure monitoring points,
which is externally determined, is maintained; a detection circuit,
wherein the detection circuit-includes a temperature sensor for
monitoring a temperature that represents the temperature of the
compartment, wherein the detection circuit produces a first
detection signal when the sensed temperature exceeds a threshold
value and a second detection signal when the sensed temperature
falls below the threshold value; and a pressure difference changer,
wherein, the pressure difference changer gradually increases the
target value of the pressure difference when the first signal is
received from the detection circuit and gradually decreases the
target value of the pressure difference when the second signal is
received from the detection circuit.
2. The controller according to claim 1, wherein the control valve
causes the displacement of the compressor to increase when the
target pressure difference is increased by the pressure difference
changer, and the control valve causes the displacement of the
compressor to decrease when the target pressure difference is
decreased by the pressure difference changer.
3. The controller according to claim 1, wherein the compressor
includes a cylinder bore, a piston reciprocally accommodated in the
cylinder, a cam plate coupled to the piston and a crank chamber for
accommodating the cam plate, wherein the crank chamber is the
control chamber.
4. The controller according to claim 1, wherein the refrigerant
circuit includes an evaporator, and wherein the temperature sensor
is located in the vicinity of the evaporator.
5. The controller according to claim 1, wherein the compressor is
driven by a vehicle engine, and wherein a controller of the engine
functions as the pressure difference changer.
6. A controller for a variable displacement compressor, wherein the
compressor is used for air conditioning a compartment and includes
a suction pressure zone, a discharge pressure zone, and a control
chamber, which is connected to the suction pressure zone and to the
discharge pressure zone, and wherein the pressure in the control
chamber is adjusted for controlling the displacement of the
compressor, the controller comprising: a refrigerant circuit
connected to the compressor, wherein two pressure monitoring points
are located in the refrigerant circuit; a control valve for
controlling the pressure in the control chamber, wherein the
control valve operates based on the actual pressure difference
between the pressure monitoring points such that a target value of
the pressure difference between the pressure monitoring points,
which is externally determined, is maintained; a detection circuit,
wherein the detection circuit includes a temperature sensor for
monitoring a temperature that represents the temperature of the
compartment, wherein the detection circuit produces a first
detection signal when the sensed temperature exceeds an upper
threshold value and a second detection signal when the sensed
temperature falls below a lower threshold value; a pressure
difference changer, wherein, the pressure difference changer
gradually increases the target value of the pressure difference
when the first signal is received from the detection circuit and
gradually decreases the target value of the pressure difference
when the second signal is received from the detection circuit.
7. The controller according to claim 6, wherein the control valve
causes the displacement of the compressor to increase when the
target pressure difference is increased by the pressure difference
changer, and the control valve causes the displacement of the
compressor to decrease when the target pressure difference is
decreased by the pressure difference changer.
8. The controller according to claim 6, wherein the compressor
includes a cylinder bore, a piston reciprocally accommodated in the
cylinder, a cam plate coupled to the piston and a crank chamber for
accommodating the cam plate, wherein the crank chamber is the
control chamber.
9. The controller according to claim 6, wherein the refrigerant
circuit includes an evaporator, and wherein the temperature sensor
is located in the vicinity of the evaporator.
10. The controller according to claim 6, wherein the compressor is
driven by a vehicle engine, and wherein a controller of the engine
functions as the pressure difference changer.
11. A controller for a variable displacement compressor, wherein
the compressor is used for air conditioning a compartment and
includes a suction pressure zone, a discharge pressure zone, and a
control chamber, which is connected to the suction pressure zone
and to the discharge pressure zone, and wherein the pressure in the
control chamber is adjusted for controlling the displacement of the
compressor, the controller comprising: a refrigerant circuit
connected to the compressor, wherein two pressure monitoring points
are located in the refrigerant circuit; a control valve for
controlling the pressure in the control chamber, wherein the
control valve operates based on the actual pressure difference
between the pressure monitoring points such that a target value of
the pressure difference between the pressure monitoring points,
which is determined externally, is maintained; a detection circuit,
wherein the detection circuit includes a temperature sensor for
monitoring a temperature that represents the temperature of the
compartment, wherein the detection circuit produces a first
detection signal when the sensed temperature exceeds an upper
threshold value and a second detection signal when the sensed
temperature falls below a lower threshold value; and a computer for
receiving the first and second detection signals and for
determining the target value of the pressure difference, wherein
the computer gradually increases the target value of the pressure
difference when the first signal is received from the detection
circuit and gradually decreases the target value of the pressure
difference when the second signal is received from the detection
circuit.
12. The controller according to claim 11, wherein the control valve
causes the displacement of the compressor to increase when the
target pressure difference is increased by the computer, and the
control valve causes the displacement of the compressor to decrease
when the target pressure difference is decreased by the
computer.
13. The controller according to claim 11, wherein the compressor
includes a cylinder bore, a piston reciprocally accommodated in the
cylinder, a cam plate coupled to the piston and a crank chamber for
accommodating the cam plate, wherein the crank chamber is the
control chamber.
14. The controller according to claim 11, wherein the refrigerant
circuit includes an evaporator, and wherein the temperature sensor
is located in the vicinity of the evaporator.
15. The controller according to claim 11, wherein the compressor is
driven by a vehicle engine, and wherein the computer controls
various engine functions.
Description
BACKGROUND OF THE INVENTION
[0001] The present invention relates to a controller of a variable
displacement compressor.
[0002] The refrigeration circuit of a typical vehicle air
conditioner includes a compressor, such as a variable displacement
swash plate type compressor. A typical variable displacement swash
plate type compressor includes a displacement control mechanism for
maintaining the pressure at the outlet of an evaporator, which will
be referred to as the suction pressure Ps, at a target value, which
will be referred to as target suction pressure. The displacement
control mechanism feedback controls the displacement of the
compressor, or the inclination angle of the swash plate, by
referring to the suction pressure Ps such that the displacement
corresponds to the cooling load. A typical displacement mechanism
includes a displacement control valve, which is called an
internally controlled valve. The internally controlled valve
detects the absolute value of the suction pressure Ps by means of a
pressure sensitive member such as a bellows or a diaphragm. The
internally controlled valve moves a valve body by the displacement
of the pressure sensing member to adjust the valve opening size.
Accordingly, the pressure in a swash plate chamber (a crank
chamber), or the crank pressure Pc is changed, which changes the
inclination of the swash plate. However, an internally controlled
valve that has a simple structure and a single target suction
pressure cannot respond to the changes in air conditioning demands.
Therefore, control valves having a target suction pressure that can
be changed by external electrical control are becoming
standard.
[0003] A typical electrically controlled control valve is a
combination of an internally controlled valve and an actuator such
as an electromagnetic solenoid, which applies an electrically
controlled force. Mechanical spring force, which acts on the
pressure sensing member is externally controlled to change the
target suction pressure. The target suction pressure is changed by
controlling a current to the electromagnetic solenoid in an analog
or a digital manner. The supplied current is controlled by a
controller having a microcomputer that is designed for air
conditioning. Specifically, the controller executes a proportional
and integral (PI) control procedure or a proportional, integral and
differential (PID) control procedure based on temperature
information from a temperature sensor located near the evaporator
or in a passenger compartment for continuously controlling the
current. As a result, the compressor theoretically maintains an
ideal displacement, or a displacement that corresponds to the
magnitude of the cooling load.
[0004] However, to execute a PI control procedure or a PID control
procedure for continuously and finely controlling the target
suction pressure, the controller, which includes a microcomputer,
must continuously receive temperature information from the
temperature sensor and compute the current supplied to a control
valve. Thus, the controller must have a high-performance
microcomputer to bear a high computation load. Even if the
controller has a high-performance microcomputer, the controller
receives temperature data relatively frequently (at an extremely
short cycle). Thus, the controller cannot be used for other
purposes, which increases the ratio of cost of the controller in
the total cost of the compressor.
[0005] In a displacement control procedure in which the absolute
value of the suction pressure Ps is used as a reference, changing
of the target suction pressure by electrical control does not
always quickly change the actual suction pressure to the target
suction pressure. This is because whether the actual suction
pressure quickly seeks a target suction pressure when the target
suction pressure is changed greatly depends on the absolute
magnitude of the cooling load. Therefore, even if the target
suction pressure is finely and continuously controlled by
controlling the current to the control valve, changes in the
compressor displacement are likely to be too slow or too
sudden.
SUMMARY OF THE INVENTION
[0006] Accordingly, it is an objective of the present invention to
provide a control device of a variable displacement compressor that
has a simple structure and improves the controllability and
response of displacement control.
[0007] To achieve the foregoing and other objectives and in
accordance with the purpose of the present invention, a controller
for a variable displacement compressor, which is used for air
conditioning a compartment, is provided. The compressor includes a
suction pressure zone, a discharge pressure zone, and a control
chamber, which is connected to the suction pressure zone and to the
discharge pressure zone. The pressure in the control chamber is
adjusted for controlling the displacement of the compressor. The
controller includes a refrigerant circuit, a control valve, a
detection circuit and a pressure difference changer. The
refrigerant circuit is connected to the compressor. Two pressure
monitoring points are located in the refrigerant circuit. The
control valve controls the pressure in the control chamber. The
control valve operates based on the actual pressure difference
between the pressure monitoring points such that a target value of
the pressure difference between the pressure monitoring points,
which is externally determined, is maintained. The detection
circuit includes a temperature sensor for monitoring a temperature
that represents the temperature of the compartment. The detection
circuit produces a first detection signal when the sensed
temperature exceeds a threshold value and a second detection signal
when the sensed temperature falls below the threshold value. The
pressure difference changer gradually increases the target value of
the pressure difference when the first signal is received from the
detection circuit and gradually decreases the target value of the
pressure difference when the second signal is received from the
detection circuit.
[0008] 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
[0009] 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:
[0010] FIG. 1 is a cross-sectional view illustrating a variable
displacement swash plate type compressor according to one
embodiment of the present invention;
[0011] FIG. 2 is a schematic diagram illustrating a refrigeration
circuit according to the embodiment of FIG. 1;
[0012] FIG. 3 is a cross-sectional view illustrating the control
valve in the compressor of FIG. 1;
[0013] FIG. 4 is a schematic cross-sectional view showing an
effective pressure receiving area of the control valve shown in
FIG. 3;
[0014] FIG. 5 is a block diagram showing a control system of the
embodiment shown in FIG. 1;
[0015] FIG. 6 is a graph showing the relationship between a
detection circuit signal and a monitored temperature;
[0016] FIG. 7 is a flowchart showing an irregular interruption
routine (1);
[0017] FIG. 8 is a flowchart showing an irregular interruption
routine (2);
[0018] FIG. 9 is a flowchart showing a regular interruption routine
(A);
[0019] FIG. 10 is a flowchart showing a regular interruption
routine (B);
[0020] FIG. 11 is a timing chart showing the relationship between a
duty ratio Dt and a detection circuit signal (a rising signal);
[0021] FIG. 12 is a timing chart showing the relationship between a
duty ratio Dt and a detection circuit signal (a falling signal);
and
[0022] FIG. 13 is a timing chart showing the relationship between a
duty ratio Dt and detection circuit signals (rising signals and
falling signals).
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0023] One embodiment according to the present invention will now
be described with reference to FIGS. 1 to 13.
[0024] As shown in FIG. 1, a variable displacement swash plate type
compressor includes a cylinder block 1, a front housing member 2,
which is secured to the front end face of the cylinder block 1, and
a rear housing member 4, which is secured to the rear end face of
the cylinder block 1. A valve plate assembly 3 is located between
the cylinder block 1 and the rear housing member 4. The cylinder
block 1, the front housing member 2, the valve plate assembly 3 and
the rear housing member 4 are secured to one another by bolts 10
(only one is shown) to form the compressor housing. 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. A crank
chamber 5 is defined between the cylinder block 1 and the front
housing member 2. A drive shaft 6 extends through the crank chamber
5 and is supported through radial bearings 8A, 8B by the housing. A
recess is formed in the center of the cylinder block 1. A coil
spring 7 and a rear thrust bearing 9B are located in the recess. A
lug plate 11 is secured to the drive shaft 6 to rotate integrally
with the drive shaft 6. A front thrust bearing 9A is located
between the lug plate 11 and the inner wall of the front housing
member 2. The drive shaft 6 is supported in the axial direction by
the rear bearing 9B, which is urged forward by the spring 7, and
the front bearing 9A.
[0025] The front end of the drive shaft 6 is connected to an
external drive source, which is a vehicle engine E in this
embodiment, through a power transmission mechanism PT. In this
embodiment, the power transmission mechanism PT is a clutchless
mechanism that includes, for example, a belt and a pulley. The
power transmission mechanism PT therefore constantly transmits
power from the engine E to the compressor when the engine E is
running. Alternatively, the mechanism PT may be a clutch mechanism
(for example, an electromagnetic clutch) that selectively transmits
power when supplied with a current.
[0026] As shown in FIG. 1, a cam plate, which is a swash plate 12
in this embodiment, is located in the crank chamber 5. A hole
extends through the middle of the swash plate 12. The drive shaft 6
extends through the hole. The swash plate 12 is connected with the
lug plate 11 and the drive shaft 6 through a coupling guide
mechanism, which is a hinge mechanism 13 in this embodiment, to
rotate integrally with the drive shaft 6. The hinge mechanism 13
includes two support arms 14 (only one is shown) and two guide pins
15 (only one is shown). Each support arm 14 projects from the rear
side of the lug plate 11. Each guide pin 15 projects from the front
side of the swash plate 12. The support arms 14 and the guide pins
15 cooperate to permit the swash plate 12 to rotate integrally with
the lug plate 11 and the drive shaft 6. Contact between the drive
shaft 6 and the wall of the swash plate center hole permits the
swash plate 12 to slide along the drive shaft 6 and to tilt with
respect to the axis of the drive shaft 6. The swash plate 12 has a
counterweight 12a located at the opposite side of the drive shaft 6
from the hinge mechanism 13.
[0027] A spring 16 is located between the lug plate 11 and the
swash plate 12. The spring 16 urges the swash plate 12 toward the
cylinder block 1, or in a direction decreasing the inclination
angle of the swash plate 12. A stopper ring 18 is fixed on the
drive shaft 6 behind the swash plate 12. A return spring 17 is
fitted about the drive shaft 6 between the stopper ring 18 and the
swash plate 12. When the inclination angle is great as shown by the
broken line in FIG. 1, the spring 17 does not apply force to the
swash plate 12. When the inclination angle is small as shown by the
solid line in FIG. 1, the spring 17 is compressed between the
stopper ring 18 and the swash plate 12 and urges the swash plate 12
away from the cylinder block 1, or in a direction increasing the
inclination angle. The normal 5 length of the spring 17 and the
location of the stopper ring 18 are determined such that the spring
17 is not fully contracted when the swash plate 12 is inclined by
the minimum inclination angle .theta.min (for example, an angle
from one to five degrees).
[0028] Several cylinder bores 1a (only one shown) are formed in the
cylinder block 1 about the drive shaft 6. The rear end of each
cylinder bore 1a is blocked by the valve plate assembly 3. A single
headed piston 20 is reciprocally accommodated in each cylinder bore
1a. Each piston 20 and the corresponding cylinder bore 1a define a
compression chamber, the volume of which is changed according to
reciprocation of the piston 20. The front portion of each piston 20
is coupled to the swash plate 12 by a pair of shoes 19. Therefore,
when the swash plate 12 rotate integrally with the drive shaft 6,
rotation of the swash plate 12 reciprocates each piston 20 by a
stroke that corresponds to the angle .theta..
[0029] A suction chamber 21, which is included in a suction
pressure zone, and discharge chamber 22, which is included in a
discharge pressure zone, are defined between the valve plate
assembly 3 and the inner wall of the rear housing member 4. The
suction chamber 21 is located approximately in the center of the
rear housing member 4, and the discharge chamber 22 surrounds the
suction chamber 21. The valve plate assembly 3 includes a suction
valve flap plate, a port plate, discharge valve flap plate and a
retainer plate. The valve plate assembly 3 has suction ports 23 and
discharge ports 25, which correspond to each cylinder bore 1a. The
valve plate assembly 3 also has suction valve flaps 24, each of
which corresponds to one of the suction ports 23, and discharge
valve flaps 26, each of which corresponds to one of the discharge
ports 25. Each cylinder bore 1a is connected to the suction chamber
21 through the corresponding suction port 23 and is connected to
the discharge chamber 22 through the corresponding discharge port
25. Refrigerant gas is drawn from the outlet of the evaporator 33
to the suction chamber 21, where the pressure is a suction pressure
Ps. When each piston 20 moves from the top dead center position to
the bottom dead center position, refrigerant gas in the suction
chamber 21 flows into the corresponding cylinder bore 1a via the
corresponding suction port 23 and suction valve flap 24. When each
piston 20 moves from the bottom dead center position to the top
dead center position, refrigerant gas in the corresponding cylinder
bore 1a is compressed to a predetermined pressure and is discharged
to the discharge chamber 22, where the pressure is a discharge
pressure Pd, via the corresponding discharge port 25 and discharge
valve 26. The highly pressurized refrigerant in the discharge
chamber 22 flows to the condenser 31.
[0030] Power from the engine E is transmitted to and rotates the
drive shaft 6. Accordingly, the swash plate 12, which is inclined
by an angle .theta., is rotated. The angle .theta. is defined by
the swash plate 12 and an imaginary plane that is perpendicular to
the drive shaft 6. Rotation of the swash plate 12 reciprocates each
piston 20 by a stroke that corresponds to the angle .theta.. As a
result, suction, compression and discharge of refrigerant gas are
repeated in the cylinder bores 1a.
[0031] The inclination angle .theta. of the swash plate 12 is
determined according to various moments acting on the swash plate
12. The moments include a rotational moment, which is based on the
centrifugal force of the rotating swash plate 12, a spring force
moment, which is based on the force of the springs 16 and 17, a
moment of inertia of the piston reciprocation, and a gas pressure
moment. The gas pressure moment is generated by the force of the
pressure in the cylinder bores 1a and the pressure in the crank
chamber 5 (crank pressure Pc). Depending on the crank pressure Pc,
the gas pressure moment acts either to increase or decrease the
inclination angle .theta. of the swash plate 12.
[0032] The gas pressure moment is adjusted by changing the crank
pressure Pc by a displacement control valve, which will be
discussed below. Accordingly, the inclination angle .theta. of the
plate 12 is adjusted to an angle between the minimum inclination
.theta.min and the maximum inclination .theta.max. Contact between
a counterweight 12a on the swash plate 12 and a stopper 11a of the
lug plate 11 prevents further inclination of the swash plate 12
from the maximum inclination .theta.max. The minimum inclination
.theta.min is determined based primarily on the forces of the
springs 16 and 17 when the gas pressure moment is maximized in the
direction in which the swash plate inclination angle .theta. is
decreased.
[0033] As described above, the crank pressure Pc is related to
changes of the inclination angle .theta. of the swash plate 12. A
mechanism for controlling the crank pressure Pc includes a bleed
passage 27, a supply passage 28 and the control valve. The passages
27, 28 are formed in the compressor housing. The bleed passage 27
connects the suction chamber 21 with the crank chamber 5. The
supply passage 28 connects the discharge chamber 22 with the crank
chamber 5. The control valve regulates the supply passage 28.
Specifically, the opening of the control valve is adjusted to
control the flow rate of highly pressurized gas supplied to the
crank chamber 5 through the supply passage 28. The crank pressure
Pc is determined by the ratio of the gas supplied to the crank
chamber 5 through the supply passage 28 and the flow rate of
refrigerant gas conducted out from the crank chamber 5 through the
bleed passage 27. As the crank pressure Pc varies, the difference
between the crank pressure Pc and the pressure in the cylinder
bores 1a varies, which changes the inclination angle .theta. of the
swash plate 12. Accordingly, the stroke of each piston 20, or the
compressor displacement, is varied.
[0034] A slight clearance (not shown) exists between the inner wall
of the each cylinder bore 1a and the corresponding piston 20. Each
clearance connects the corresponding cylinder bore 1a with the
crank chamber 5. The discharge pressure zone, which is connected to
the crank chamber 5, includes the discharge chamber 22 and one or
more of the cylinder bores 1a in which the piston 20 is in the
compression stroke. When each piston 20 compresses the gas in the
associated cylinder bore 1a, some of the refrigerant gas in the
cylinder bore 1a leaks into the crank chamber 5 through the
clearance between the cylinder bore 1a and the piston 20. The
leaking gas is referred to as blowby gas. The blowby gas increases
the pressure of the crank chamber 5. The discharge pressure zone in
this embodiment includes the discharge chamber 22 and the cylinder
bores 1a.
[0035] As shown in FIGS. 1 and 2, a refrigeration circuit, or a
refrigerant circuit, of a vehicle air conditioner includes the
variable displacement swash plate type compressor and an external
refrigerant circuit 30. The external refrigerant circuit 30
includes, for example, a condenser 31, a decompression device and
an evaporator 33. The decompression device is an expansion valve 32
in this embodiment. The opening of the expansion valve 32 is
feedback-controlled based on the temperature detected by a heat
sensitive tube 34 at the outlet of the evaporator 33 and the
evaporation pressure, or the pressure at the evaporator outlet. The
expansion valve 32 supplies liquid refrigerant to the evaporator 33
to regulate the flow rate in the external refrigerant circuit 30.
The amount of the supplied refrigerant corresponds to the thermal
load. A downstream pipe 35 is located in a downstream portion of
the refrigerant circuit 30 to connect the outlet of the evaporator
33 to the suction chamber 21 of the compressor. An upstream pipe 36
is located in an upstream portion of the refrigerant circuit 30 to
connect the discharge chamber 22 of the compressor to the inlet of
the condenser 31. The compressor draws refrigerant gas from the
downstream portion of the refrigeration circuit 30 and compresses
the gas. The compressor then discharges the compressed gas to the
discharge chamber 22, which is connected to the upstream portion of
the circuit 30.
[0036] The greater the displacement of the compressor is, the
higher the flow rate of refrigerant in the refrigeration circuit
is. The greater the flow rate of the refrigerant is, the greater
the pressure loss per unit length of the circuit is. That is, the
pressure loss between two points in the refrigeration circuit
corresponds to the flow rate of refrigerant in the circuit.
Detecting the pressure difference .DELTA.P(t) between two points
P1, P2 permits the displacement of the compressor to be indirectly
detected. In this embodiment, two pressure monitoring points P1, P2
are defined in the upstream pipe 36. The first pressure monitoring
point P1 is located in the discharge chamber 22, which is the most
upstream section of the upstream pipe 36. The second pressure
monitoring point P2 is located in the upstream pipe 36 and is
spaced from the first point P1 by a predetermined distance. A part
of the control valve is exposed to the pressure PdH, or the
discharge pressure Pd, at the first point P1 by a first pressure
introduction passage 37. Another part of the control valve is
exposed to a pressure PdL at the second point P2 by a second
pressure introduction passage 38. The control valve feedback
process uses the pressure difference expressed by
.DELTA.P(t)=PdH-PdL to estimate the compressor displacement and to
feedback control the displacement.
[0037] The displacement control valve shown in FIG. 3 mechanically
detects the pressure difference between the pressure monitoring
points P1, P2 and adjusts the valve opening based on the detected
pressure difference.
[0038] As shown in FIG. 3, the control valve includes an inlet
valve and a solenoid. The inlet valve is arranged in an upper
portion of the valve, while the solenoid is arranged in a lower
portion of the valve. The inlet valve adjusts the opening size
(throttle amount) of the supply passage 28, which connects the
discharge chamber 22 to the crank chamber 5. The solenoid is an
electromagnetic actuator for urging a rod 40 located in the control
valve based on current supplied from an outside source. The
solenoid functions as an actuator 100 for changing a target
pressure difference.
[0039] The rod 40 includes a distal portion 41, a coupler portion
42 and a proximal guide portion 44. The guide portion 44 includes a
valve body 43, which is located in the center of the rod 40. The
diameter of the distal portion 41, the coupler portion 42 and the
guide portion 44 are represented by d1, d2 and d3, respectively.
The diameters satisfy the inequality d2<d1<d3. The
cross-sectional area SB of the distal portion 41 is represented by
.pi.(d1/2).sup.2. The cross-sectional area SC of the coupler
portion 42 is represented by .pi.(d2/2).sup.2. The cross-sectional
area SD of the guide portion 44 is represented by
.pi.(d3/2).sup.2.
[0040] The control valve has a valve housing 45. The housing 45
includes a cap 45a and an upper portion 45b and a lower portion
45c. The cap 45a is fixed to the end of the upper portion 45b. The
upper portion 45b defines the shape of the inlet valve portion. The
lower portion 45c defines the shape of the solenoid. A valve
chamber 46 and a communication passage 47 are formed in the upper
portion 45b. A pressure sensing chamber 48 is defined between the
upper portion 45b and the cap 45a.
[0041] The rod 40 extends through the valve chamber 46, the
communication passage 47 an the pressure sensing chamber 48. The
rod 40 moves axially, or in the vertical direction as viewed in the
drawing. The valve chamber 46 is connected to the communication
passage 47 depending on the position of the rod 40. The
communication passage 47 is disconnected from the pressure sensing
chamber 48 by a wall, which is a part of the valve housing 45. A
guide hole 49 is formed in the wall to receive the rod 40. The
diameter of the guide hole 49 is equal to the diameter d1 of the
distal portion 41. The communication passage 47 is axially aligned
with the guide hole 49, and the diameter of the communication
passage 47 is equal to the diameter dl of the distal portion 41.
That is, the area of the communication passage 47 and the area of
the guide hole 49 are equal to the area SB of the distal portion
41.
[0042] The bottom of the valve chamber 46 is formed by the upper
surface of a fixed iron core 62. A Pd port 51 extends radially from
the valve chamber 46. The valve chamber 46 is connected to the
discharge chamber 22 through the Pd port 51 and the upstream
section of the supply passage 28. A Pc port 52 radially extends
from the communication passage 47. The communication passage 47 is
connected to the crank chamber 5 through the downstream section of
the supply passage 28 and the Pc port 52. Therefore, the Pd port
51, the valve chamber 46, the communication passage 47 and the Pc
port 52 are formed in the control valve and form a part of the
supply passage 28, which connects the discharge chamber 22 with the
crank chamber 5.
[0043] The valve body 43 of the rod 40 is located in the valve
chamber 46. The diameter d1 of the communication passage 47 is
greater than the diameter d2 of the coupler portion 42 and smaller
than the diameter d3 of the guide portion 44. Thus, a step is
formed between the valve chamber 46 and the communication passage
47. The step functions as a valve seat 53, and the communication
passage 47 functions as a valve hole. When the rod 40 is moved from
the position of FIG. 3, or the lowermost position, to the uppermost
position, at which the valve body 43 contacts the valve seat 53,
the communication passage 47 is disconnected from the valve chamber
46. That is, the valve body 43 is an inlet valve body that controls
the opening size of the supply passage 28.
[0044] A movable wall 54 is located in the pressure sensing chamber
48. The movable wall 54 divides the pressure sensing chamber 48
into a first pressure chamber 55 and a second pressure chamber 56.
The movable wall 54 does not permit fluid to move between the first
pressure chamber 55 and the second pressure chamber 56. The
cross-sectional area SA of the movable wall 54 is greater than the
cross-sectional area SB of the guide hole 49 (SB<SA).
[0045] The first pressure chamber 55 is constantly connected to the
discharge chamber 22, which is the upstream pressure monitoring
point P1, by a P1 port 55a formed in the cap 45a and the first
passage 37.
[0046] The second pressure chamber 56 is constantly connected to
the second pressure monitoring point P2 through a P2 port 56a
formed in the upper portion 45b and the second passage 38. The
first pressure chamber 55 is exposed to the discharge pressure Pd,
which is the pressure PdH. The second pressure chamber 56 is
exposed to the pressure PdL at the second pressure monitoring point
P2. The upper side of the movable wall 54 receives the pressure PdH
and the lower side receives the pressure PdL. The distal portion 41
of the rod 40 is located in the second pressure chamber 56. The
distal end of the distal portion 41 is coupled to the movable wall
54. A spring 57 is located in the second pressure chamber 56. The
spring 57 urges the movable wall 54 toward the first pressure
chamber 55.
[0047] The solenoid (the actuator 100 for changing the target
pressure difference) includes a cup-shaped cylinder 61, which is
fixed in the lower portion 45c. A stationary iron core 62 is fitted
into an upper opening of the cylinder 61. The stationary core 62
defines a solenoid chamber 63 in the cylinder 61. A movable iron
core 64 is located in the solenoid chamber 63. The movable iron
core 64 is moved axially. The stationary core 62 has a guide hole
65 through which the guide portion 44 extends. There is a clearance
(not shown) between the guide hole 65 and the guide portion 44. The
clearance communicates the valve chamber 46 with the solenoid
chamber 63. Thus, the solenoid chamber 63 is exposed to the
discharge pressure Pd, to which the valve chamber 46 is
exposed.
[0048] The proximal portion of the rod 40 is located in the
solenoid chamber 63. The lower end of the guide portion 44 is
fitted into a hole formed in the center of the movable iron core
64. The movable iron core 64 is crimped to the guide portion 44.
Thus, the movable core 64 moves integrally with the rod 40. A
spring 66 is located between the stationary core 62 and the movable
core 64. The spring 66 urges the movable core 64 and the rod 40
downward such that the movable core 64 moves away from stationary
core 62.
[0049] A coil 67 is wound about the stationary core 62 and the
movable core 64. The coil 67 receives drive signals from a drive
circuit 72 based on commands from an ECU 70 for the engine E. The
coil 67 generates an electromagnetic force F that corresponds to
the value of the current from the drive circuit 72. The
electromagnetic force F urges the movable core 64 toward the
stationary core 62, which lifts the rod 40. The current to the coil
67 may be varied in an analog fashion. Alternatively, the current
may be duty controlled, that is, the duty ratio Dt of the current
may be controlled. In this case, a greater duty ratio Dt represents
a smaller opening size of the control valve and a smaller duty
ratio Dt represents a greater opening size of the control
valve.
[0050] The opening size of the control valve is determined by the
position of the rod 40. The rod 40 has the valve body 43, which
functions as an inlet valve body. Forces acting on several parts of
the rod 40 will now be explained to describe the operating
conditions and the characteristics of the control valve.
[0051] The upper surface of the distal portion 41 receives a
downward force, which is the resultant of the force fl of the
spring 57 and the pressures acting on the upper and the lower sides
of the movable wall 54. The pressure receiving area on the upper
side of the wall 54 is represented by SA. The pressure receiving
area of the lower side of the wall 54 is represented by (SA-SB).
The pressure receiving area of the lower end of the distal portion
41 is represented by (SB-SC). The crank pressure Pc applies an
upward force to the lower end of the distal portion 41. Assume
downward forces have positive values. The sum .SIGMA.F1 of the
forces acting on the distal portion 41 is represented by the
following equation.
.SIGMA.F1=PdH.multidot.SA-PdL(SA-SB)-f1-Pc(SB-SC) Equation I
[0052] A downward force f2 of the spring 66 and an upward
electromagnetic force F act on the guide portion 44, which includes
the valve body portion 43.
[0053] The pressures that act on the exposed surfaces of the valve
body 43, the guide portion 44 and the movable iron core 64 will now
be described with reference to FIG. 4. The pressures are simplified
as follows. First, the upper end surface of the valve body 43 is
divided into the inside section and the outside section by an
imaginary cylinder, which is shown by broken lines in FIG. 4. The
imaginary cylinder corresponds to the wall of the communication
passage 47. The crank pressure Pc acts in a downward direction on
the inside section (area: SB-SC). The discharge pressure Pd acts in
a downward direction on the outside section (area: SD-SB). Taking
the pressure balance between the upper and lower surfaces of the
movable iron core 64 into account, the discharge pressure Pd, to
which the solenoid chamber 63 is exposed, acts on the area
corresponding to the cross-sectional area SD of the guide portion
44 to urge the guide portion 44 upward. If the total force
.SIGMA.F2 that acts on the valve body 43 and the guide portion 44,
defining the upward direction as the positive direction, are
summed, .SIGMA.F2 is expressed by the following equation.
.SIGMA.F2=F-f2-Pc(SB-SC)-Pd(SD-SB)+Pd.multidot.SD=F-f2-Pc(SB-SC)+Pd.multid-
ot.SB Equation II
[0054] In the process of calculating equation II, -Pc.multidot.SD
was canceled by +Pc.multidot.SD, and the term Pc.multidot.SB
remained. That is, if the net force based on the discharge pressure
Pd that acts on the upper and lower surfaces of the guide portion
44 is viewed as a force that acts on the lower surface of the guide
portion 44, the effective pressure receiving area of the guide
portion 44 regarding the discharge pressure Pd is equal to the area
SB (SB=SD-(SD-SB)). As far as the discharge pressure Pd is
concerned, the effective pressure receiving area of the guide
portion 44 is equal to the cross-sectional area SB of the
communication passage 47 regardless of the cross-sectional area SD
of the guide portion 44. When pressures of the same kind act on
both ends of a member such as a rod, the pressure receiving area
having an effect that is not canceled is called the effective
pressure receiving surface area.
[0055] Since the rod 40 is an integrated member formed by
connecting the guide portion 44 to the distal portion 41 with
coupler portion 42, its position is determined by the physical
balance of .SIGMA.F1=.SIGMA.F2. In the equation
.SIGMA.F1=.SIGMA.F2, the terms Pc(SB-SC) can be canceled. As a
result, the following equation III is obtained.
(PdH-PdL)SA-Pd.multidot.SB+PdL.multidot.SB=F+f1-f2 Equation III
[0056] Since the first pressure monitoring point P1 is located in
the discharge chamber 22, the pressure Pd is equal to the pressure
PdH (Pd=PdH). If Pd is replaced by PdH, equation III is converted
into the following equations IV and V.
(PdH-PdL)SA-(PdH-PdL)SB=F+f1-f2 Equation IV
PdH-PdL=(F+f1-f2)/(SA-SB) Equation V
[0057] In equation V, f1, f2, SA and SB are fixed parameters that
are primarily defined in the steps of mechanical design, and the
electromagnetic force F is a variable parameter that changes in
accordance with the power supplied to the coil 67.
[0058] As apparent from equation V, the pressure difference
.DELTA.P(t) (.DELTA.P(t)=PdH-PdL), is determined only by duty
controlling the current supplied to the coil 67. That is, a target
value TPD of the pressure difference is adjusted by externally
controlling the control valve.
[0059] Equation V contains no pressure parameters such as the crank
pressure Pc and the discharge pressure Pd, other than the pressure
difference expressed by PdH-PdL. Thus, the crank pressure Pc and
the discharge pressure Pd do not influence the position of the rod
40. In other words, pressure parameters other than the pressure
difference do not affect the movement of the rod 40, and the
control valve is regulated based only on the pressure difference
.DELTA.P(t), the electromagnetic force F and the spring forces f1,
f2.
[0060] The opening size of the control valve is determined in the
following manner. When no current is supplied to the coil 67, or
when the duty ratio Dt is zero percent, the spring 66 positions the
rod 40 at the lowest position shown in FIG. 3. The valve body 43 is
spaced from the valve seat 53 by the greatest distance, which fully
opens the control valve. When a current of the minimum duty ratio
is supplied to the coil 67, the upward electromagnetic force F is
greater than the downward force f2 of the spring 66. The net upward
force (F-f2) generated by the solenoid and the spring 66 acts
against the net downward force of the pressure difference (PdH-PdL)
and the spring 57. As a result, the position of the valve body 43
relative to the valve seat 53 is determined such that equation V is
satisfied, which determines the opening size of the control
valve.
[0061] Accordingly, the flow rate of gas to the crank chamber 5
through the supply passage 28 is determined. Then, the crank
pressure Pc is adjusted in accordance with the relationship between
the flow rate of gas through the supply passage 28 and the flow
rate of gas flowing out from the crank chamber 5 through the bleed
passage 27. That is, controlling the opening size of the control
valve controls the crank pressure Pc. When the electromagnetic
force F is constant, the control valve functions as a constant flow
rate valve and is actuated based on the target pressure difference
TPD, which corresponds to the electromagnetic force F. However,
since electromagnetic force F can be externally changed to adjust
the target pressure difference TPD, the control valve can vary the
displacement of the compressor.
[0062] Control System
[0063] As shown in FIGS. 2, 3 and 5, the control valve is connected
to a pressure difference changer, which is an engine ECU 70 in this
embodiment, through the drive circuit 72. The engine ECU 70 mainly
controls the engine E. As shown in FIG. 5, the ECU 70 includes a
CPU, a ROM, a RAM, a timer and an input-output interface circuit.
The ROM stores various control programs (see flowcharts of FIGS. 7
to 10) and initial data. The RAM has a working memory area. The
timer generates clock pulse signals by either hardware or software.
The clock pulse signals are at least used as regular interruption
signals for notifying the CPU of the starting time of regular
interruption routines. The input-output interface circuit has input
and output terminals. An external information detection apparatus
71 is connected to input terminals. The drive circuit 72 is
connected to output terminals. The engine ECU 70 computes an
appropriate duty ratio Dt based on the information from the
apparatus 71 and commands the drive circuit 72 to output a drive
signal having the computed duty ratio Dt. The drive circuit 72
outputs the instructed drive signal having the duty ratio Dt to the
coil 67 of the control valve. The electromagnetic force F of the
solenoid is determined according to the duty ratio Dt. Accordingly,
the opening size of the control valve is continuously adjusted,
which quickly changes the crank pressure Pc and the stroke of each
piston 20. The piston stroke represents the compressor displacement
and the torque.
[0064] The external information detection apparatus 71 includes
various sensors. The sensors of the detection apparatus 71 may
include, for example, an A/C switch 81, a vehicle speed sensor 82,
an engine speed sensor 83, a throttle sensor (or an acceleration
pedal sensor) 84 and a detection circuit 85. The A/C switch 81 is
an ON/OFF switch of the air conditioner operated by a passenger.
The A/C switch 81 provides the engine ECU 70 with information
regarding the ON/OFF state of the air conditioner. The vehicle
speed sensor 82 and the engine speed sensor 83 provide the engine
ECU 70 with information regarding the vehicle speed V and the
engine speed NE. The throttle sensor 84 detects the inclination
angle, or the opening size, of a throttle valve located in the
intake passage of the engine. The throttle opening size represents
the degree of depression Ac(t) of the acceleration pedal in the
vehicle.
[0065] The detection circuit 85 is located in the vicinity of the
evaporator 33 (see FIG. 2) and provides the engine ECU 70 with
information regarding the temperature in the vicinity of the
evaporator 33. The temperature information will be referred to as a
detection circuit signal. The temperature in the vicinity of the
evaporator 33 corresponds to the temperature of the surface of the
evaporator 33 and to the temperature of the passenger compartment.
The detection circuit 85 includes a temperature sensor, which is a
thermistor 86 in this embodiment, for monitoring the temperature in
the vicinity of the evaporator 33 and a signal output circuit 87
for generating and outputting the detection circuit signal based on
changes of the resistance of the thermistor 86.
[0066] The signal output circuit 87 compares the monitored
temperature with threshold temperatures. When the monitored
temperature falls below one of the threshold temperatures or
surpasses another, the circuit 87 outputs the detection circuit
signal. FIG. 6 shows the relationship between the monitored
temperature and the detection circuit signal. The threshold
temperatures are a lower limit temperature T1 (for example, three
degrees centigrade) and an upper limit temperature T2 (for example,
four degrees centigrade). The monitored temperature rises due to
changes in the relationship between the flow rate of the
refrigerant in the evaporator and the compartment temperature. When
the monitored temperature surpasses the upper limit temperature T2,
the signal output circuit 87 outputs an ON signal (a rising
signal).
[0067] When the monitored temperature falls below the lower limit
temperature T1, the signal output circuit 87 outputs an OFF signal
(falling signal). Since the determination values differ when the
signal is switched from OFF to ON from when the signal is switched
from ON to OFF, there is a hysteresis. The threshold temperatures,
which are three degrees centigrade and four degrees centigrade in
this embodiment, are determined such that air sent to the passenger
compartment is sufficiently cooled without forming frost the
evaporator. Frost on the evaporator reduces the cooling
efficiency.
[0068] A controller of the compressor at least includes the engine
ECU 70, the detection circuit 85 and the control valve.
[0069] Duty control procedure by the ECU 70 will be described with
reference to flowcharts and timing charts (FIGS. 7 to 13). The ECU
70 normally controls the engine E by, for example, controlling the
fuel supply amount. In addition, the ECU 70 regularly and
irregularly performs interruptions for controlling the air
conditioner.
[0070] FIG. 7 is a flowchart of an irregular interruption routine
(1), which is executed for starting and stopping air conditioning.
When the A/C switch 81 is turned on or off and a signal
representing the switching reaches the engine ECU 70, the ECU 70
judges that there is an interrupt request. In this case, the ECU 70
stops controlling the engine E and starts the irregular
interruption routine (1).
[0071] If the A/C switch 81 is switched from OFF to ON in step S71,
the ECU 70 moves to step S72. In step S72, the ECU 70 initializes
the duty ratio Dt. That is, the ECU 70 sets the duty ratio Dt to an
initial value Dt(ini), which is, for example, fifty percent. The
opening size of the control valve corresponds to the initial duty
ratio Dt(ini). The crank pressure Pc is changed accordingly and the
compressor displacement is set to a predetermined initial
level.
[0072] If the A/C switch 81 is switched from ON to OFF in step S71,
the ECU 70 moves to step S73. In step S73, the ECU 70 sets the duty
ratio Dt to zero, which maximizes the opening size of the control
valve. Accordingly, the crank pressure Pc is quickly increased and
the inclination angle .theta. is minimized. The compressor
displacement is thus minimized. After either steps S72, S73, the
ECU 70 terminates the interruption and starts controlling the
engine E again.
[0073] FIG. 8 is a flowchart of an irregular interruption routine
(2), which is executed when the A/C switch is on. When the signal
from the detection circuit 85 changes, the engine ECU 70 judges
that there is an interruption request. In this case, the ECU 70
stops controlling the engine E and starts the irregular
interruption routine (2). If the ECU 70 receives a rising signal in
step S81, the ECU 70 moves to step S82. In step S82, the ECU 70
starts regular interruption routine (A), which is shown in FIG. 9.
If the ECU 70 receives a falling signal in step S81, the ECU 70
moves to step S83. In step S83, the ECU 70 starts a regular
interruption routine (B), which is shown in FIG. 10. After
executing either steps S82 and S83, the ECU 70 terminates the
interruption routine (2) and starts controlling the engine E
again.
[0074] When the duty ratio Dt is the initial value Dt(ini), the
compressor displacement is changed, which lowers the temperature in
the vicinity of the evaporator 33. When the monitored temperature
falls below the lower limit temperature T1, the ECU 70 receives a
falling signal from the detection circuit 85 and thus starts the
routine (B). The ECU 70 regularly repeats the routine (B) until the
ECU 70 receives a rising signal and starts the routine (A). The
routine (B) is executed in synchronization with clock signals from
the timer.
[0075] When the engine ECU 70 stops controlling the engine E and
starts the routine (B), the ECU 70 decreases the current duty ratio
Dt by an amount .DELTA.D in step S101. A decrease in the duty ratio
Dt represents a decrease of the target pressure difference TPD and
a decrease of the refrigerant flow rate or a decrease in the
compressor displacement. Accordingly, the air conditioning is
controlled to lessen cooling.
[0076] In step S102, the ECU 70 judges whether the current duty
ratio Dt, which was computed by subtracting the amount .DELTA.D
from the previous duty ratio Dt, is smaller than a predetermined
lower limit value Dt(min). If the outcome of step S102 is negative,
the current duty ratio Dt is greater than the lower limit value
Dt(min). In this case, the ECU 70 moves to step S103 and commands
the drive circuit 72 to change the duty ratio Dt, which slightly
weakens the electromagnetic force F. Accordingly, the target
pressure difference TPD is slightly lowered.
[0077] Then, since balance between the forces on the rod 40 is not
achieved with the current pressure difference .DELTA.P(t), the rod
40 is moved downward, which reduces the force applied by the spring
66. Thus, the reduced downward force f2 of the return spring 66 is
countered by the reduced upward electromagnetic force F, and the
valve body 43 is positioned such that equation V is satisfied
again. As a result, the opening size of the control valve, that is,
the opening size of the supply passage 28, is increased, which
increases the crank pressure Pc. Accordingly, the difference
between the crank pressure Pc and the pressure of the cylinder
bores 1a increases, and the inclination angle .theta. of the swash
plate 12 is decreased. Accordingly, the compressor displacement is
decreased. When the discharge displacement of the compressor is
decreased, the heat reduction performance of the evaporator 33 is
also reduced, the passenger compartment temperature, or the
monitored temperature, is increased, and the pressure difference
between the points P1 and P2 is decreased.
[0078] If the outcome of step S102 is positive, the ECU 70 sets the
duty ratio Dt to the lower limit value Dt(min) in step S104 and
commands the drive circuit 72 to operate at the lower limit value
Dt(min) in step S103. The lower limit value Dt(min) may be
zero.
[0079] As the routine (B) is repeated, the duty ratio Dt, or the
target pressure difference TPD, is gradually decreased. The timing
chart of FIG. 12 shows changes of the duty ratio Dt when the
routine (B) is repeated. When receiving a falling signal from the
detection circuit 85, the ECU 70 keeps gradually decreasing the
duty ratio Dt by the amount .DELTA.D at a time in synchronization
with the timer clock until the ECU 70 receives a rising signal.
Accordingly, the duty ratio Dt is gradually decreased to the lower
limit value Dt(min) (see the graph of Dt from t3 to t4 in FIG. 12).
Then, as long as the ECU 70 does not receive a rising signal from
the detection circuit 85, the duty ratio Dt is maintained at the
lower limit value Dt(min) (see t4 and after in the graph of FIG.
12).
[0080] A decrease in the duty ratio Dt decreases the compressor
displacement and reduces the heat reduction performance of the
evaporator 33. Accordingly, the compartment temperature, or the
monitored temperature, is gradually increased. When the monitored
temperature surpasses the upper limit temperature T2, the engine
ECU 70 receives a rising signal from the detection circuit 85. The
ECU 70 then repeats the regular interruption routine (A), which is
shown in FIG. 9, until the ECU 70 receives a falling signal.
[0081] When the engine ECU 70 stops controlling the engine E and
starts the routine (A), the ECU 70 increases the duty ratio Dt by
the amount .DELTA.D in step S91. An increase in the duty ratio Dt
increases the target pressure difference TPD, which increases the
refrigerant flow rate and the compressor displacement. Accordingly,
the cooling performance is increased.
[0082] In step S92, the ECU 70 judges whether the current duty
ratio Dt, which was computed by adding the amount .DELTA.D to the
previous duty ratio Dt, is greater than a predetermined upper limit
value Dt(max). If the outcome of step S92 is negative, the current
duty ratio Dt is smaller than the upper limit value Dt(max). In
this case, the ECU 70 moves to step S93 and commands the drive
circuit 72 to change the duty ratio Dt, which slightly strengthen,
the electromagnetic force F. Accordingly, the target pressure
difference TPD is slightly raised.
[0083] Then, since balance between the forces on the rod 40 is not
achieved with the current pressure difference .DELTA.P(t), the rod
40 is moved upward, which increases the force applied by the spring
66. Thus, the increased downward force f2 of the return spring 66
is countered by the increased upward electromagnetic force F, and
the valve body 43 is positioned such that equation V is satisfied
again. As a result, the opening size of the control valve, that is,
the opening size of the supply passage 28, is decreased, which
decreases the crank pressure Pc. Accordingly, the difference
between the crank pressure Pc and the pressure of the cylinder
bores 1a decreases, and the inclination angle .theta. of the swash
plate 12 is increased. Accordingly, the compressor displacement is
increased. When the discharge displacement of the compressor is
increased, the heat reduction performance of the evaporator 33 is
also increased, the passenger compartment temperature, or the
monitored temperature, is decreased, and the pressure difference
between the points P1 and P2 is increased.
[0084] If the outcome of step S92 is positive, the ECU 70 sets the
duty ratio Dt to the upper limit value Dt(max) in step S94 and
commands the drive circuit 72 to operate at the upper limit value
Dt(max) in step S93. As the routine (A) is repeated, the duty ratio
Dt, or the target pressure difference TPD, is gradually increased.
The timing chart of FIG. 11 shows changes of the duty ratio Dt when
the routine (A) is repeated. When receiving a rising signal from
the detection circuit 85, the ECU 70 gradually increases the duty
ratio Dt by the amount .DELTA.D at a time in synchronization with
the timer clock until the ECU 70 receives a falling signal.
Accordingly, the duty ratio Dt is gradually increased to the upper
limit value Dt(min) (see the graph of Dt from t1 to t2 in FIG. 11).
Then, as long as the ECU 70 does not receive a rising signal from
the detection circuit 85, the duty ratio Dt is maintained at the
upper limit value Dt(max) (see t2 and after in the graph of FIG.
11).
[0085] An increase in the duty ratio Dt increases the compressor
displacement and increases the heat reduction performance of the
evaporator 33. Accordingly, the compartment temperature, or the
monitored temperature, is gradually decreased. When the monitored
temperature falls below the lower limit temperature T1, the ECU 70
then repeats the regular interruption routine (B), which is shown
in FIG. 10, until the ECU 70 receives a rising signal.
[0086] The engine ECU 70 continues to gradually increase or
decrease the duty ratio Dt, or the target suction pressure TPD,
until the ECU 70 receives a signal (a detection circuit signal)
that indicates the monitored temperature crosses one of the
threshold temperatures from the detection circuit 85. When
receiving such a signal, the ECU 70 reverse the changing direction
of the target pressure difference TPD. Thus, the target pressure
difference TPD (duty ratio Dt) is alternately increased and
decreased.
[0087] If there is no abrupt changes of thermal load, the increases
and decreases of the duty ratio Dt, the duty ratio Dt changes along
line 131 in the timing chart of FIG. 13 from a macroscopic
viewpoint. Changes of the monitored temperature, or increases and
decreases of the monitored temperature between the threshold
temperatures T1 and T2, the detection circuit 85 alternately
outputs rising signals and falling signals. Every time the circuit
85 switches between a rising signal and a falling signal, the duty
ratio Dt repeats increases and decreases with a constant amplitude
above and below a center value DtMid(t). The center value DtMid(t)
may be variable or constant. For example, a dashed line 132
represents the center value DtMid(t). In other words, while the
engine ECU 70 changes the detection circuit signal between On and
OFF in a binary fashion, which maintains the duty ratio Dt in the
vicinity of the center value DtMid(t) in a certain amplitude.
[0088] As described above, the duty ratio Dt, or the pressure
difference TPD, is quickly adjusted when the thermal load on the
evaporator 33 is changed. The flow rate of refrigerant is adjusted
accordingly, which maintains the temperature in the vicinity of the
evaporator 33 at a temperature suitable for cooling the passenger
compartment.
[0089] The illustrated embodiment has the following advantages.
[0090] The temperature in the vicinity of the evaporator 33 is
maintained at a level suitable for cooling by a simple procedure.
That is, the ECU 70 simply increases or decreases in response to a
rising signal or a falling signal from the detection circuit 85. In
other words, the procedures for optimizing the temperature in the
vicinity of the evaporator 33 are sufficiently simple to be
performed as interruptions by the ECU 70, which reduces the
calculation load on the ECU 70. Thus, there is no need for an
expensive controller specialized for air conditioning, and the
engine ECU 70, which is mainly used for controlling the engine E,
is used for air conditioning.
[0091] In the illustrated embodiment, the threshold temperatures,
which are compared with the temperature monitored by the detection
circuit 85, include the lower and upper limit temperatures T1, T2.
Also, there is a hysteresis in which the temperature at which a
rising signal is generated is different from the temperature at
which a falling signal is generated. If there is only one threshold
temperature, hunting may occur. Compared to a system having a
single threshold temperature, the illustrated embodiment stably
controls the compressor displacement without applying an excessive
load on the compressor. Hunting of the detection circuit 85 refers
to a case where the monitored temperature surpasses and falls below
a single threshold temperature and the resulting detection circuit
signals are excessively generated during a short time.
[0092] The suction pressure Ps is greatly influenced by changes in
the thermal load on the evaporator 33. In the illustrated
embodiment, the suction pressure Ps is not directly referred to for
controlling the opening size of the displacement control valve.
Instead, the pressure difference .DELTA.P(t)(.DELTA.P(t)=PdH-PdL)
between the two pressure monitoring points P1 and P2 is directly
controlled for feedback controlling the compressor
displacement.
[0093] Therefore, the compressor displacement is quickly controlled
from the outside without being influenced by the thermal load on
the evaporator 33.
[0094] The control valve shown in FIG. 3 functions as an internally
controlled valve. Specifically, as long as the electromagnetic
force F is constant, the control valve shown in FIG. 3 maintains
the target pressure difference TPD, which is determined by the
forces F, f1, f2 and the areas SA, SB, and automatically controls
the compressor displacement to a level that corresponds to the
target pressure difference TPD. The electromagnetic force F can be
externally changed for changing the target pressure difference TPD.
The compressor displacement is changed accordingly.
[0095] 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.
[0096] The thermistor 86 and the signal output circuit 87 in the
detection circuit 85 may be integrated or separated. If the
thermistor 86 and the circuit 87 are separated, the thermistor 86
needs to monitor a temperature, which is the temperature of the
evaporator 33 in the illustrated embodiment.
[0097] The upper limit value Dt(max) and the lower limit value
Dt(min) of the duty ratio Dt, which are used in steps S94 and S104,
need not be used.
[0098] The upper limit temperature T2 and the lower limit
temperature T1 may be replaced by a single threshold
temperature.
[0099] In the illustrated embodiment, the engine ECU 70 functions
as the target pressure difference changer However, the target
pressure difference TPD may be changed by a separate controller.
Compared to PI control and PID control, in which the target
pressure difference is continuously and finely controlled, the
control procedure of the illustrated embodiment is simple, which
reduces the cost of the controller.
[0100] In the illustrated embodiment, the present invention is
applied to a reciprocal piston type compressor. However, the
present invention may be applied to rotary compressors such as a
variable displacement scroll type compressor disclosed in Japanese
Unexamined Patent Publication No. 11-324930.
[0101] In the illustrated embodiment, the upstream pressure
monitoring point P1 is located in the discharge chamber 22, and the
downstream pressure monitoring point P2 is located in the upstream
pipe 36. However, the upstream pressure monitoring point P1 may be
located in the downstream pipe 35 and the downstream pressure
monitoring point P2 may be located in the suction chamber 21.
Alternatively, the upstream pressure monitoring point P1 may be
located either in the discharge chamber or the upstream pipe 36 and
the downstream pressure monitoring point P2 may be located either
in the suction chamber 21 or the downstream pipe 35. Also, the
upstream pressure monitoring point P1 may be located either in the
discharge chamber 22 and the upstream pipe 36 and the downstream
pressure monitoring point P2 may be located in the crank chamber 5.
Further, the upstream pressure monitoring point P1 may be located
in the crank chamber 5 and the downstream pressure monitoring point
P2 may be located either in the suction chamber 21 or the
downstream pipe 35.
[0102] 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.
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