U.S. patent application number 09/777523 was filed with the patent office on 2001-10-11 for controller for variable displacement compressor.
Invention is credited to Adaniya, Taku, Kawaguchi, Masahiro, Kimura, Kazuya, Matsubara, Ryo, Ota, Masaki, Suitou, Ken.
Application Number | 20010027658 09/777523 |
Document ID | / |
Family ID | 18554796 |
Filed Date | 2001-10-11 |
United States Patent
Application |
20010027658 |
Kind Code |
A1 |
Ota, Masaki ; et
al. |
October 11, 2001 |
Controller for variable displacement compressor
Abstract
A controller controls a displacement of a compressor which is
installed in a refrigerant circuit of an air-conditioning system.
The controller has a pressure difference detector. An air
conditioning switch turns the air conditioning system on. A
temperature detector detects the temperature in a compartment. A
control valve controls the displacement of the compressor such that
the pressure difference detected by the pressure difference
detector approaches a target value. A computer determines the
target value. The computer changes the target value, depending on
the detected temperature. The computer limits the target value to a
target value limit. After a predetermined time expires from when
the air conditioning switch is turned on, the computer changes the
limit. This permits the displacement of the compressor to be
promptly and reliably changed.
Inventors: |
Ota, Masaki; (Kariya-shi,
JP) ; Kimura, Kazuya; (Kariya-shi, JP) ;
Kawaguchi, Masahiro; (Kariya-shi, JP) ; Suitou,
Ken; (Kariya-shi, JP) ; Matsubara, Ryo;
(Kariya-shi, JP) ; Adaniya, Taku; (Kariya-shi,
JP) |
Correspondence
Address: |
Kurt E. Richter
MORGAN & FINNEGAN, L.L.P.
345 Park Avenue
New York
NY
10154
US
|
Family ID: |
18554796 |
Appl. No.: |
09/777523 |
Filed: |
February 6, 2001 |
Current U.S.
Class: |
62/228.3 ;
417/222.2 |
Current CPC
Class: |
F04B 27/1804 20130101;
F04B 2027/1813 20130101; F04B 2027/1827 20130101; F04B 2207/03
20130101; F04B 2205/07 20130101; F04B 49/065 20130101; F04B
2027/1854 20130101; F04B 2027/1859 20130101 |
Class at
Publication: |
62/228.3 ;
417/222.2 |
International
Class: |
F04B 001/26 |
Foreign Application Data
Date |
Code |
Application Number |
Feb 7, 2000 |
JP |
2000-029551 |
Claims
What is claimed is:
1. A controller for controlling the displacement of a compressor,
which is included in a refrigeration circuit of an air conditioning
system, the controller comprising: a pressure difference detector
for detecting the pressure difference between two pressure
monitoring points located in the refrigeration circuit; an air
conditioning switch for turning the air conditioning system on; a
temperature detector for detecting the temperature in a
compartment; a mechanism for controlling the displacement of the
compressor such that the pressure difference detected by the
pressure difference detector approaches a target value; and a
computer for determining the target value, wherein the computer
changes the target value, depending on the detected temperature,
and limits the target value to a target value limit, and after a
predetermined time expires from when the air conditioning switch is
turned on, the computer changes the limit.
2. The controller according to claim 1, wherein the temperature
detector has a temperature sensor and a signal output circuit,
wherein the signal output circuit sends a rising signal, which
indicates that the temperature of the compartment is increasing,
and a falling signal, which indicates that the compartment
temperature is falling, to the computer, and the computer increases
the target value when one of the rising or falling signal is
received and decreases the target value when the other of the
rising or falling signal is received.
3. The controller according to claim 1, wherein the temperature
detector has a temperature sensor and a signal output circuit,
wherein the signal output circuit sends a rising signal, which
indicates that the temperature of the compartment has increased
above a first level, and a falling signal, which indicates that the
compartment temperature has fallen below a second level, to the
computer, and the computer increases the target value when one of
the rising or falling signal is received and decreases the target
value when the other of the rising or falling signal is
received.
4. The controller according to claim 1, wherein the refrigerant
circuit includes an evaporator, wherein the temperature detector is
located near the evaporator.
5. The controller according to claim 1, wherein the compressor
includes a drive plate and a crank chamber for accommodating the
drive plate, wherein the inclination angle of the drive plate
changes in accordance with the pressure in the crank chamber to
vary the displacement of the compressor, and wherein the mechanism
for controlling the displacement comprises: a control valve, the
opening size of which is changed in accordance with an external
command to adjust the pressure in the crank chamber, wherein the
control valve includes the pressure difference detector, and the
control valve changes the opening size of the control valve based
on the detected pressure difference.
6. The controller according to claim 5, wherein the control valve
comprises: a valve body, the position of which is changed by a
force produced by the pressure difference detector; and an
actuator, wherein the actuator applies a force based on the target
value to the valve body according to the external command.
7. The controller according to claim 1, wherein the refrigerant
circuit includes an evaporator, the compressor has a discharge
pressure zone, and the pressure monitoring points are located in
the refrigerant circuit between the evaporator and the discharge
pressure zone.
8. A method for controlling the displacement of a compressor
installed in a refrigerant circuit of an air-conditioning system,
the method including: detecting the pressure difference between two
pressure monitoring points located in the refrigerant circuit by a
pressure difference detector; detecting the temperature in a
compartment; determining a target value of the pressure difference,
wherein the determining includes limiting the target value to a
target value limit; changing the target value depending on the
detected temperature; changing the target value limit after a
predetermined time expires from when the air conditioning system is
activated; and controlling the displacement of the compressor such
that the pressure difference detected by the pressure difference
detector approaches the target value.
Description
BACKGROUND OF THE INVENTION
[0001] The present invention relates to a controller for variable
displacement compressors to control displacement.
[0002] A refrigeration circuit of a typical vehicle
air-conditioning system includes a condenser, an expansion valve,
which functions as a depressurizing device, an evaporator and a
compressor. The compressor draws refrigerant gas from the
evaporator and compresses the gas. The compressor then discharges
the gas to the condenser. The evaporator transfers heat between the
refrigerant in the circuit and air in the passenger compartment.
Heat from air that flows about the evaporator is transferred to the
refrigerant flowing through the evaporator in accordance with the
thermal load or the cooling load. The pressure of the refrigerant
gas at the outlet of the evaporator represents the magnitude of the
thermal load.
[0003] A vehicle variable displacement swash plate type compressor
has a displacement control mechanism for setting the pressure
(suction pressure Ps) in the vicinity of the outlet of the
evaporator to a predetermined target suction pressure. The
mechanism adjusts the compressor displacement by changing the
inclination angle of the swash plate such that the flow rate of
refrigerant corresponds to the cooling load. The displacement
control mechanism has a control valve. The control valve includes a
pressure sensing member, which is a bellows or a diaphragm. The
suction pressure is detected by the Ps pressure sensing member. A
valve opening is adjusted in accordance with the displacement of
the pressure sensing member, which changes the pressure in a crank
chamber, or crank pressure Pc.
[0004] A simple control valve that imposes a single target suction
pressure cannot control the air conditioning performance
accurately. Therefore, an electromagnetic control valve that
changes a target suction pressure in accordance with an external
current has been proposed. Such a control valve includes an
electromagnetic actuator such as a solenoid. The actuator changes a
force acting on the pressure sensing member in accordance with an
external current to adjust the target suction pressure.
[0005] According to the above-described control method, however,
even if the target suction pressure is changed by electric control,
the actual suction pressure may not reach the target suction
pressure spontaneously. That is, the cooling load is likely to
affect whether or not the actual suction pressure well responds to
a change in the target suction pressure. It is not therefore
possible to promptly and reliably alter the displacement of a
compressor even if the actual suction pressure is regulated as
needed by electric control.
SUMMARY OF THE INVENTION
[0006] Accordingly, it is an object of the present invention to
provide a control apparatus for a variable displacement type
compressor, which can promptly and reliably change the displacement
of the compressor.
[0007] To achieve the above objective, the present invention
provides a controller for controlling the displacement of a
compressor. The compressor is included in a refrigeration circuit
of an air conditioning system. The controller comprises a pressure
difference detector. The pressure difference detector detects the
pressure difference between two pressure monitoring points located
in the refrigeration circuit. An air conditioning switch turns the
air conditioning system on. A temperature detector detects the
temperature in a compartment. A mechanism controls the displacement
of the compressor such that the pressure difference detected by the
pressure difference detector approaches a target value. A computer
determines the target value. The computer changes the target value,
depending on the detected temperature. The computer limits the
target value to a target value limit. After a predetermined time
expires from when the air conditioning switch is turned on, the
computer changes the limit.
[0008] To achieve the above objective, the present invention also
provides a method for controlling the displacement of a compressor
installed in a refrigerant circuit of an air-conditioning system.
The method comprises detecting the pressure difference between two
pressure monitoring points located in the refrigerant circuit,
detecting the temperature in a compartment, determining a target
value of the pressure difference. The determining includes limiting
the target value to a target value limit, changing the target value
depending on the detected temperature, changing the target value
limit after a predetermined time expires from when the air
conditioning system is activated, and controlling the displacement
of the compressor such that the pressure difference detected by the
pressure difference detector approaches the target value.
[0009] 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
[0010] The features of the present invention that are believed to
be novel are set forth with particularity in the appended claims.
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:
[0011] FIG. 1 is a cross-sectional diagram of a variable
displacement type swash plate compressor according to one
embodiment of the present invention;
[0012] FIG. 2 is a circuit diagram illustrating the outline of a
cooling circuit according to the embodiment;
[0013] FIG. 3 is a cross-sectional diagram of a control valve
provided in the compressor in FIG. 1;
[0014] FIG. 4 is a partly enlarged cross-sectional diagram for
explaining the positioning of an actuation rod;
[0015] FIG. 5 is a block diagram showing the electric structure of
a control apparatus for the compressor in FIG. 1;
[0016] FIG. 6 is a graph showing the correlation between a
detection circuit signal and a monitored temperature;
[0017] FIG. 7 is a flowchart of an irregular interruption routine
(1);
[0018] FIG. 8 is a flowchart of a regular interruption routine
(C);
[0019] FIG. 9 is a flowchart of an irregular interruption routine
(2);
[0020] FIG. 10 is a flowchart of a regular interruption routine
(A);
[0021] FIG. 11 is a flowchart of a regular interruption routine
(B);
[0022] FIG. 12 is a time chart showing the correlation between a
duty ratio and a detection circuit signal;
[0023] FIG. 13 is a time chart showing the correlation between the
duty ratio and the detection circuit signal; and
[0024] FIG. 14 is a time chart showing the correlation between the
duty ratio and the detection circuit signal.
DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0025] A vehicle air-conditioning system according to a first
embodiment of the present invention will now be described with
reference to FIGS. 1 to 14.
[0026] The compressor shown in FIG. 1 is a swash plate type
variable displacement reciprocal compressor. The 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 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 3 and the rear housing
member 4 are secured to one another by bolts 10 (only one is shown)
to form the compressor housing.
[0027] 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 rotatably supported through radial bearings
8A, 8B by the housing. A recess is formed in the center of the
cylinder block 1. A spring 7 and a rear thrust bearing 9B are
located in the recess. A lug plate 11 is secured to the drive shaft
6 in the crank chamber 5 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. A rear thrust
bearing 9B is located adjacent to the rear end of the drive shaft
6. The drive shaft 6 is supported in the axial direction by the
rear bearing 9B, which is urged in a forward direction by the
spring 7, and the front bearing 9A.
[0028] The front end of the drive shaft 6 is connected to an
external drive source, which is an 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. Alternatively, the
mechanism PT may be a clutch mechanism (for example, an
electromagnetic clutch) that selectively transmits power in
accordance with the value of an externally supplied current.
[0029] A drive plate, which is a swash plate 12 in this embodiment,
is accommodated in the crank chamber 5. The swash plate 12 has a
hole formed in the center. The drive shaft 6 extends through the
hole in the swash plate 12. The swash plate 12 is coupled to the
lug plate 11 by a guide mechanism, which is a hinge mechanism 13 in
this embodiment. 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 swash plate 12. The swash
plate 12 rotates integrally with the lug plate 11 and drive shaft
6. The swash plate 12 slides along the drive shaft 6 and tilts with
respect to the axis of the drive shaft 6. The swash plate 12 has a
counterweight 12a located at the opposite side of drive shaft 6
with respect to the drive hinge mechanism 13.
[0030] A first spring 16 is located between the lug plate 11 and
the swash plate 12 to reduce the angle of the swash plate 12. The
first spring 16 urges the swash plate 12 toward the cylinder block
1, or in the direction decreasing the inclination of the swash
plate 12. The inclination of the swash plate 12 is defined by an
inclination angle, which is the angle between the swash plate 12
and a plane perpendicular to the drive shaft 6. A stopper ring 18
is fixed on the drive shaft 6 behind the swash plate 12. A second
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 broken line in FIG. 1, the second spring 17 does not
apply force to the swash plate 12 and other members. When the
inclination angle is small, as shown by solid lines in FIG. 1, the
second 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 length of the second spring 17 and the location of the
stopper ring 18 are determined such that the second spring 17 is
not fully contracted when the swash plate 12 is inclined by the
minimum inclination angle (for example, an angle from one to five
degrees).
[0031] Cylinder bores la (only one shown) are formed in the
cylinder block 1. The cylinder bores 1a are arranged at equal
angular intervals about the drive shaft 6. The rear end of each
cylinder bore 1a is blocked by the valve plate 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,
rotation of the swash plate 12 reciprocates each piston 20 by a
stroke that corresponds to the angle of the swash plate 12.
[0032] A suction chamber 21 and a discharge chamber 22 are defined
between the valve plate 3 and the rear housing member 4. The
discharge chamber 22 surrounds the suction chamber 21. The valve
plate 3 has suction ports 23 and discharge ports 25, which
correspond to each cylinder bore 1a. The valve plate 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. The suction ports 23
connect the suction chamber 21 with the cylinder bores 1a. The
discharge ports 25 connect the cylinder bores 1a with the discharge
chamber 22.
[0033] When each piston 20 moves from the top dead center position
to the bottom dead center position, refrigerant gas in the suction
chamber 21, which is a suction pressure zone, flows into the
corresponding cylinder bore 1a via the corresponding suction port
23 and suction valve 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, which is a discharge pressure zone, via the corresponding
discharge port 25 and discharge valve 26.
[0034] Power of the engine E is transmitted to and rotates the
drive shaft 6. Accordingly, the swash plate 12, which is inclined
by an angle, is rotated. Rotation of the swash plate 12
reciprocates each piston 20 by a stroke that corresponds to the
angle. As a result, suction, compression and discharge of
refrigerant gas are repeated in the cylinder bores 1a.
[0035] The inclination angle 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). The gas pressure moment is adjusted
by changing the crank pressure Pc by a displacement control valve
CV, which will be discussed below. Accordingly, the inclination
angle of the plate 12 is adjusted to an angle between the maximum
inclination and the minimum inclination. The contact between the
counterweight 12a and a stopper 11a of the lug plate 11 prevents
further inclination of the swash plate 12 from the maximum
inclination. The minimum inclination is determined based chiefly on
the forces of the springs 16 and 17 when the gas pressure moment is
maximized in the direction by which the swash plate inclination is
decreased.
[0036] A mechanism for controlling the crank pressure Pc includes a
bleeding passage 27, a supply passage 28 and the control valve CV
as shown in FIGS. 1 and 2. The passages 27, 28 are formed in the
housing. The bleeding 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 CV
is located in the supply passage 28.
[0037] The control valve CV changes the opening of the supply
passage 28 to adjust the flow rate of refrigerant gas from the
discharge chamber 22 to the crank chamber 5. The crank pressure Pc
is changed in accordance with the relationship between the flow
rate of refrigerant gas from the discharge chamber 22 to the crank
chamber 5 and the flow rate of refrigerant gas flowing out from the
crank chamber 5 to the suction chamber 21 through the bleeding
passage 27. The difference between the crank pressure Pc and the
pressure in the cylinder bores 1a is changed in accordance with the
crank pressure Pc, which varies the inclination angle of the swash
plate 12. This alters the stroke of each piston 20 and the
compressor displacement.
[0038] FIG. 1 illustrates a refrigeration circuit of a vehicle
air-conditioning system. The refrigeration circuit has a compressor
and an external refrigeration circuit 30. The refrigeration circuit
30 includes, for example, a condenser 31, an expansion valve 32 and
an evaporator 33. 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. The expansion
valve 32 supplies refrigerant, the amount of which corresponds to
the thermal load on the evaporator 33, to regulate the flow rate. A
passage 35 is provided in a downstream portion of the external
refrigerant circuit 30 for connecting the outlet of the evaporator
33 to the suction chamber 21 of the compressor. A passage 36 is
provided in an upstream portion of the external refrigerant circuit
30 for connecting 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 upstream portion of the circuit 30.
[0039] 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, the greater the
pressure loss per unit length in 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. By
detecting the pressure difference .DELTA.P(t) (.DELTA.P(t)=PsH-PsL)
between two points P1, P2, the displacement of the compressor is
detected indirectly. In this embodiment, the point P1 is located in
the discharge chamber 22 and is an upstream pressure monitoring
point. The point P2 is located in the passage 36 at a position
spaced from the point 1 by a predetermined distance and is a
downstream pressure monitoring point. The gas pressure PdH at the
point P1 is applied to the displacement control valve CV through a
first pressure detecting passage 37. The gas pressure PdL at the
point P2 is applied to the displacement control valve CV through a
second pressure detecting passage 38. The displacement control
valve CV performs a feedback control procedure for the compressor
displacement in accordance with the pressure difference between the
point P1 and the point P2 (PdH-PdL).
[0040] A displacement control valve CV shown in FIG. 3 mechanically
detects a differential pressure .DELTA.P(t) between two points in a
cooling circuit, and directly uses the differential pressure to
adjust its opening degree.
[0041] The control valve CV includes an inlet valve portion and a
solenoid. The inlet valve portion adjusts the opening size of the
supply passage 28 connecting the discharge chamber 22 to the crank
chamber 5. The solenoid functions as an electromagnetic actuator
100 that controls a rod 40 provided in the control valve CV in
accordance with an external electric current supply. A pressure
difference receiving portion 41 is provided at a distal end of the
rod 40. A valve body 43 is provided at a substantially intermediate
portion of the rod 40. The pressure difference receiving portion 41
is connected to the valve body 43 by a connecting portion 42. The
rod 40 further includes a guide portion 44. The valve body 43 forms
part of the guide portion 44. The diameter d1 of the pressure
difference receiving portion 41, the diameter d2 of the connecting
portion 42, and the diameter d3 of the guide portion 44 (the valve
body 43) satisfy the following condition: d2<d1<d3. The
cross-sectional area SB of the pressure difference receiving
portion 41 in a plane perpendicular to the axis of the rod 40 is
.pi.(d1/2).sup.2. The cross-sectional area SC of the connecting
portion 42 in a plane perpendicular to the axis of the rod 40 is
.pi.(d2/2).sup.2. The cross-sectional area SD of the guide portion
44 (the valve body 43) in a plane perpendicular to the axis of the
rod 40 is .pi.(d3/2).sup.2.
[0042] The control valve CV has a valve housing 45 including a cap
45a, an upper body section 45b, and a lower body section 45c, as
shown in FIG. 3. A valve chamber 46 and a communication passage 47
are formed in the upper body section 45b. A pressure sensing
chamber 48 is provided between the upper body section 45b and the
cap 45a.
[0043] The rod 40 extends through the valve chamber 46, the
communication passage 47, and the pressure sensing chamber 48 and
moves along the axis of the control valve CV. The valve chamber 46
is selectively connected to and disconnected from the passage 47 in
accordance with the position of the rod 40. The communication
passage 47 is completely blocked from the pressure sensing chamber
48 by a wall that formes part of the valve housing 45. The diameter
of the passage 47 and the diameter of a guide hole 49 are equal to
the diameter d1 of the pressure difference receiving portion 41 of
the rod 40.
[0044] The bottom of the valve chamber 46 is formed by the upper
surface of a fixed iron core 62. An inlet port 51 extends radially
from the valve chamber 46. The valve chamber 46 is connected to the
discharge chamber 22 through the inlet port 51 and the upstream
portion of the supply passage 28. An outlet port 52 radially
extends from the communication passage 47. The communication
passage 47 is connected to the crank chamber 5 through the
downstream portion of the supply passage 28 and the outlet port 52.
Therefore, the inlet port 51, the valve chamber 46, the
communication passage 47 and the outlet port 52, which are formed
in the control valve CV, form a part of the supply passage 28,
which connects the discharge chamber 22 with the crank chamber
5.
[0045] The valve body 43 of the rod 40 is located in the valve
chamber 46. The diameter d1 of the communication passage 47 is
larger than the diameter d2 of the connecting portion 42 of the rod
40 and is smaller than the diameter d3 of the large diameter the
end portion 44. A valve seat 53 is formed about the opening of the
communication passage 47, which functions as a valve hole. If the
rod 40 is moved from the position shown in FIG. 3, or its lowest
position, to its highest position, where the valve body 43 contacts
the valve seat 53, the communication passage 47 is closed. That is,
the valve body 43 of the rod 40 functions as an inlet valve body,
which controls the opening size of the supply passage 28. In this
description, upward is the direction in which the rod 40 closes the
communication passage 47, and downward is the direction in which
the rod 40 opens the passage 47.
[0046] An axially movable wall 54, or partition member, is provided
in the pressure sensing chamber 48. The movable wall 54 axially
divides the pressure sensing chamber 48 into two sections, or a
first pressure chamber 55 and a second pressure chamber 56. The
movable wall 54 separates the first pressure chamber 55 from the
second pressure chamber 56. The first pressure chamber 55 is thus
isolated from the second pressure chamber 56. The cross-sectional
area SA of the movable wall 54 in a plane perpendicular to the axis
of the rod 40 is greater than the cross-sectional area SB of the
passage 47 or the guide hole 49 (SB<SA) perpendicular to the
axis of the rod 40.
[0047] The first pressure chamber 55 is constantly connected to the
discharge chamber 22, in which the point P1 is located, through a
P1 port 55a formed in the cap 45a and the first pressure detecting
passage 37. The second pressure chamber 56 is constantly connected
to the point P2 through a P2 port 56a, which extends through the
upper body section 45b, and the second pressure detecting passage
38. Accordingly, the discharge pressure Pd is applied to the first
pressure chamber 55 as the pressure PdH, and the pressure PdL at
the point P2 located in the passage 36 is applied to the second
pressure chamber 56. That is, the upper side of the movable wall 54
is exposed to the pressure PdH, and the lower side of the movable
wall 54 is exposed to the pressure PdL, as viewed in FIG. 3. A
distal end, or upper end, of the pressure difference receiving
portion 41 of the rod 40 is located in the second pressure chamber
56. The movable wall 54 is secured to the distal end of the
pressure difference receiving portion 41. A dampener spring 57 is
provided in the second pressure chamber 56 for urging the movable
wall 54 toward the first pressure chamber 55.
[0048] The solenoid 100, or the electromagnetic actuator 100, which
controls the rod 40 in accordance with an external electric current
supply, has an accommodating cylinder 61 with a closed end. The
fixed iron core 62 is fitted in an upper section of the cylinder
61, and a solenoid chamber 63 is formed in the cylinder 61. The
solenoid chamber 63 accommodates a movable iron core 64, or
plunger. The movable core 64 axially moves in the solenoid chamber
63.
[0049] A guide hole 65 extends axially through the middle of the
fixed core 62. The guide hole 65 accommodates the guide portion 44
of the rod 40. The guide portion 44 axially moves in the guide hole
65. A clearance (not shown) is defined between the wall of the
guide hole 65 and the guide portion 44. The clearance connects the
valve chamber 46 to the solenoid chamber 63. The solenoid chamber
63 thus receives the discharge pressure Pd, like the valve chamber
46.
[0050] A lower end of the guide portion 44, or the proximal end of
the rod 40, is fitted in a hole formed in the middle of the movable
core 64 and is fixed to the movable core 64. The movable core 64
thus moves integrally with the rod 40. A return spring 66 is
provided between the fixed core 62 and the movable core 64. The
return spring 66 urges the movable core 64 in a direction to
separate the movable core 64 from the fixed core 62, or downward.
That is, the return spring 66 functions as an initializing means
that returns the movable core 64 and the rod 40 to their lowermost
positions.
[0051] A coil 67 is wound around the fixed core 62 and the movable
core 64. A drive circuit 72 sends a drive signal indicating a
predetermined duty ratio Dt to the coil 67, in accordance with an
instruction of a controller 70. The coil 67 then generates
electromagnetic force F that corresponds to the duty ratio Dt or in
accordance with an external electric current supply to the coil 67.
The electromagnetic force F attracts the movable core 64 toward the
fixed core 62, thus moving the rod 40 upward. The electric current
supply to the coil 67 may be controlled by an analog electric
current control procedure, a duty control procedure, in which the
duty ratio Dt is altered as necessary, or a pulse width modulation
control procedure (PWM control procedure). As the duty ratio Dt
becomes smaller, the opening size of the control valve CV becomes
larger. That is, as the duty ratio Dt becomes larger, the opening
size of the control valve CV becomes smaller.
[0052] The opening size of the control valve CV of FIG. 3 is
determined in accordance with the position of the rod 40 including
the valve body 43. The operational conditions and characteristics
of the control valve CV are determined in relation to the forces
acting on various portions of the rod 40.
[0053] An upper side of the pressure difference receiving portion
41 of the rod 40 receives a downward force that is generated in
accordance with the equilibrium of the pressure difference between
the first pressure chamber 55 and the second pressure chamber 56
(PdH-PdL) and the upward force f1 of the dampener spring 57. The
pressure receiving area of the upper side of the movable wall 54 is
SA, and the pressure receiving area of the lower side of the
movable wall 54 is SA-SB. Further, an upward force that is
generated by the crank pressure Pc is applied to the lower side of
the pressure difference receiving portion 41, the pressure
receiving area of which is SB-SC. If the downward direction is
considered to be the positive direction, a total force .SIGMA.F1
applied to the pressure difference receiving portion 41 is
represented by the following equation (1):
.SIGMA.F1=PdH.multidot.SA-PdL(SA-SB)-f1-Pc(SB-SC) (1)
[0054] The guide portion 44 of the rod 40 receives an upward force
that is generated in accordance with the equilibrium between the
electromagnetic force F of the coil 67 and the downward force f2 of
the return spring 66.
[0055] The pressures acting on the valve body 43, the guide portion
44, and the movable core 64 will now be explained with reference to
FIG. 13. The upper side of the valve body 43 is divided into two
sections, an inner section and an outer section, with respect to a
hypothetical cylindrical surface extending around the axis of the
rod 40 and along the wall of the communication passage 47 (as
indicated by broken lines in FIG. 13). As shown in FIG. 4, the
crank pressure Pc applies an axially downward force over a
cross-sectional area SB-SC of the inner section, and the discharge
pressure Pd applies an axially downward force over a
cross-sectional area SD-SB of the outer section. Further, the
discharge pressure Pd applies an upward axial force to a lower side
of the guide portion 44 over a cross-sectional area SD in a plane
perpendicular to the axis of the guide portion 44. If the upward
direction is considered to be the positive direction, the total
force .SIGMA.F applied to the valve body 43 and the guide portion
44 is indicated by the following equation (2):
.SIGMA.F2=F-f2-Pc(SB-SC)-Pd(SD-SB)+Pd.multidot.SD=F-f2-Pc(SB-SC)+Pd.multid-
ot.SB (2)
[0056] If it is assumed that the discharge pressure Pd is applied
only to the lower side of the guide portion 44 of the rod 40,
equation (2) indicates that the effective pressure receiving area
of the rod 40 is represented by the following equation:
SD-(SD-SB)=SB. That is, the effective pressure receiving area of
the guide portion 44, which receives the discharge pressure Pd,
corresponds to the cross-sectional area SB of the passage 47,
regardless of the cross-sectional area SD of the guide portion 44.
When opposite ends of a rod or the like receive the same type of
pressure, the difference between the opposed surfaces areas that
receive the pressure is defined as the effective pressure receiving
area.
[0057] Equation (2) is satisfied even if the cross-sectional area
of the valve body 43 and that of the guide portion 44 is SB and the
valve body 43 is inserted in the passage 47 (the cross-sectional
area of which is SB), and if the crank pressure Pc acts on the
upper side of the valve body 43 and the discharge pressure Pd is
applied to the lower side of the guide portion 44.
[0058] The rod 40 is formed by the pressure difference receiving
portion 41 and the guide portion 44 that are connected by the
connecting portion 42. The rod 40 is thus positioned to satisfy the
following condition: .SIGMA.F1=.SIGMA.F2. Based on equations (1),
(2), the following equation (3) is obtained:
(PdH-PdL)SA-Pd.multidot.SB+PdL.multidot.SB=F-f2+f1 (3)
[0059] In this embodiment, the point P1 is located in the discharge
chamber 22. Accordingly, the following equation is satisfied:
Pd=PdH. If this equation is applied to equation (3), equations (4),
(5) are obtained.
(PdH-PdL)SA-(PdH-PdL)SB=F-f2+f1 (4)
PdH-PdL=(F-f2+f1)/(SA-SB) (5)
[0060] In equation (5), only the electromagnetic force F is varied
in accordance with an electric current supplied to the coil 67. The
opening size of the displacement control valve CV shown in FIG. 3
is adjusted by performing an external duty control procedure for
the coil 67 to alter a target value for the pressure difference
between P1 and P2, or .DELTA.P(t)=PdH-PdL (which is a target
pressure difference TPD). In other words, the control valve CV is
externally controlled to alter the target pressure difference TPD.
A target pressure difference determining means of the control valve
CV shown in FIG. 3 is formed by the electronic actuator 100, the
return spring 66, and the dampener spring 57.
[0061] Equation (5) does not have pressure parameters (values
including Pc or Pd) other than the pressure difference between P1
and P2 (PdH-PdL) This indicates that the rod 40 is positioned
regardless of the crank pressure Pc and the discharge pressure Pd.
In other words, the rod 40 is positioned regardless of pressure
parameters other than the pressure difference between P1 and P2.
The control valve CV of FIG. 3 is thus smoothly operated only in
relation to the equilibrium of the force caused by the pressure
difference between P1 and P2 .DELTA.P(t), the electromagnetic force
F, and the urging forces f1, f2.
[0062] The operational characteristics of the displacement control
valve of the first embodiment will hereafter be described. When the
current supply to the coil 67 is null (Dt=0%), the return spring 66
maintains the rod 40 at its lowermost position, as shown in FIG. 3.
In this state, the valve body 43 of the rod 40 is spaced from the
valve seat 53 by a maximum distance. The inlet valve portion of the
control valve CV is thus completely opened. If an electric current
with a minimum duty ratio Dt is supplied to the coil 67, the upward
electromagnetic force F becomes greater than the downward force f2
of the spring 66. The upward force (F-f2) matches a downward force
determined by the equilibrium between the pressure difference
between P1 and P2 (PdH-PdL) and the force f1 of the dampener spring
57. Accordingly, the valve body 43 is positioned with respect to
the valve seat 53 to satisfy the equation (5), thus determining the
opening size of the control valve CV. This determines the amount of
gas flowing to the crank chamber 5 through the supply passage 28.
The opening size of the CV does not regulate the passage 27. The
crank pressure Pc is thus adjusted.
[0063] As long as the electromagnetic force F is constant, the
control valve CV of FIG. 3 is operated with a target pressure
difference TPD corresponding to the current electromagnetic force
F. If the electromagnetic force F is altered in accordance with an
external electric current supply, the control valve CV changes the
target pressure difference TPD accordingly.
[0064] As shown in FIGS. 2, 3 and 5, an air-conditioner for a
vehicle comprises a controller 70. The controller 70 that is
connected to the control valve CV via a drive circuit 72. As shown
in FIG. 5, the controller 70 is a control unit comprising a CPU, a
ROM, a RAM, a clock signal generator, a counter and an input/output
(I/O) interface circuit.
[0065] Various control programs (see flowcharts in FIGS. 7 to 11)
and initial data, which will be discussed later, are stored in the
ROM. The RAM provides a work memory area. The clock signal
generator generates a clock pulse signal at a predetermined
interval. The clock signal is used as a regular interruption signal
to inform the CPU of the timing of initiating a regular
interruption routine. The counter counts the pulses of the clock
signal generated by the clock signal generator.
[0066] The I/O interface circuit of the controller 70 is provided
with a plurality of input and output terminals. An external
information detector 71 is connected to the input terminals of the
I/O interface circuit, and the drive circuit 72 is connected to the
output terminals of the I/O interface circuit.
[0067] The controller 70 computes a proper duty ratio Dt based on
various external information provided from the external information
detector 71 and sends a signal representing the duty ratio Dt to
the drive circuit 72. The drive circuit 72 sends a drive signal
having the duty ratio Dt to a coil 67 of the control valve CV. The
electromagnetic force F of the solenoid section changes in
accordance with the duty ratio Dt of the drive signal supplied to
the coil 67. This ensures arbitrary real-time adjustment of the
degree of opening of the control valve CV to promptly change the
crank pressure Pc and the piston stroke. In other words, the
displacement of the compressor is swiftly altered.
[0068] Sensors that constitute the external information detector 71
include an air conditioner switch (hereinafter referred to as "A/C
switch") 81, a vehicle speed sensor 82, an engine speed sensor 83,
a gas pedal sensor 84 which detects the angle or the degree of
opening of the throttle valve, and a detection circuit 85. The A/C
switch 81 switches ON or OFF the air conditioner and is manipulated
by a passenger in a vehicle. The A/C switch 81 provides the
controller 70 with information about the ON/OFF setting of the air
conditioner. The vehicle speed sensor 82 and engine speed sensor 83
provide the controller 70 with information about a vehicle speed V
and engine speed NE. The gas pedal sensor 84 detects the angle (or
the degree of opening) of the throttle valve in the intake passage
of the engine E. The throttle valve angle (the degree of throttle
opening) reflects the amount of depression of the acceleration
pedal by the driver of the vehicle (i.e., an acceleration position
Ac(t)).
[0069] The detection circuit 85, which serves as temperature
detector, is provided near an evaporator 33 (see FIG. 2). The
detection circuit 85 provides the controller 70 with information
about the temperature near the evaporator 33. The temperature near
the evaporator 33 is correlated with the surface temperature of the
evaporator 33 and the passenger compartment temperature that is the
target in the air-conditioning control. The detection circuit 85
has a thermistor 86 as a temperature sensor to detect the
temperature near the evaporator 33 and a signal output circuit 87.
The signal output circuit 87 outputs a detection circuit signal
based on a change in the resistance of the thermistor 86,
corresponds to a change in the temperature.
[0070] The signal output circuit 87 compares the detected
temperature with a predetermined threshold temperature. When the
level relationship between those two temperatures is reversed, the
signal output circuit 87 sends out a detection circuit signal
indicating that event. FIG. 6 shows the correlation between the
monitored temperature and the detection circuit signal. Set in the
signal output circuit 87 are a lower temperature limit T1 (e.g.,
3.degree. C.) and an upper temperature limit T2 (e.g., 4.degree.
C.) as threshold values. The signal output circuit 87 outputs an ON
signal the instant the monitored temperature rises above the upper
temperature limit T2 due to the balance between the amount of the
coolant flowing in the evaporator 33 and the temperature of the
passenger compartment.
[0071] The signal output circuit 87 outputs an OFF signal the
instant the monitored temperature shifts below the lower
temperature limit T1. That is, the threshold value for switching
from the OFF state to the ON state differs from the threshold value
for switching from the ON state to the OFF state. The reason for
using two different threshold values is to avoid hunting, which is
likely to occur when a single threshold value is used, by a
so-called hysteresis determination pattern. The threshold
temperatures of, for example, 3.degree. C. and 4.degree. C. in this
embodiment are adequate temperatures to avoid the frosting of the
surface of the evaporator 33 and to produce cool air to cool the
passenger compartment of the vehicle.
[0072] The controller 70, the external information detector 71, the
drive circuit 72 and the control valve CV constitute the control
apparatus for the compressor.
[0073] Referring now to the flowcharts in FIGS. 7 to 11 and the
time charts of FIGS. 12-14, the duty control of the control valve
CV by the controller 70 will be described below. The controller 70
normally performs engine control on the engine E including, for
example, control of the supply of fuel. The controller 70 also
performs regular and irregular interruption processes associated
with air-conditioning.
[0074] The flowchart in FIG. 7 shows an irregular interruption
routine (1) concerning interruption to start and stop the
air-conditioning process. When the A/C switch 81 is manipulated and
a switch signal is sent to the controller 70, the controller 70
interrupts the engine control and initiates irregular interruption
routine 1.
[0075] In step S71, the controller 70 determines the ON/OFF state
of the A/C switch 81. When the A/C switch 81 is switched ON, the
controller 70 performs various kinds of initialization in step S72.
For example, the controller 70 sets the duty ratio Dt of the drive
signal to be supplied to the control valve CV via the drive circuit
72 to an initial value DtIni (e.g., DtIni-50%), resets the counter
and sets the limit value (upper limit value) DtMax(x) of the duty
ratio Dt to DtMax(100) (duty ratio of 100%). The initialization
causes the degree of opening of the control valve CV to correspond
to the initial value DtIni of the duty ratio Dt, which changes Pc
to a level that corresponds to the opening size of the control
valve CV. As a result, the displacement of the compressor is
controlled to a predetermined initial displacement that corresponds
to the change in the crank pressure Pc. The setting of the upper
limit value DtMax(x) of the duty ratio Dt to DtMax(100) allows the
electromagnetic force F or the target pressure difference TPD to be
changed within a range up to the maximum value allowed by the
structure of the control valve CV.
[0076] When the A/C switch 81 is switched OFF, the controller 70
sets the duty ratio Dt to zero in step S73. When the duty ratio Dt
is zero, the control valve CV is open to the maximum degree. This
increases the crank pressure Pc and swiftly minimizes the
inclination angle of the swash plate 12, thereby minimizing the
displacement of the compressor. After step S72 or S73, the
interruption routine is terminated and the controller 70 restarts
engine control.
[0077] The flowchart in FIG. 8 shows regular interruption routine
C, which is effective when the A/C switch 81 is ON. The controller
70 interrupts engine control in synchronization with a clock signal
from the clock signal generator and executes regular interruption
routine C. In step S61, the controller 70 determines whether the
counter has counted the pulses of the clock signal up to a
predetermined value or not. In other words, the controller 70
determines whether or not a predetermined time (e.g., 5 to 20
minutes) has elapsed since the A/C switch 81 was switched ON.
[0078] When the determination is NO in step S61, the controller 70
terminates regular interruption routine C while the upper limit
value DtMax(x) kept at DtMax(100) and restarts engine control. When
the determination is YES in step S61, the controller 70 changes the
upper limit value DtMax(x) of the duty ratio Dt to, for example, 60
to 75% (DtMax(70) (duty ratio of 70%) in this embodiment) from
DtMax(100) in step S62. That is, after the predetermined time has
elapsed since the switching of the A/C switch 81 to the ON state,
the controller 70 changes the upper limit of the duty ratio, which
determines the electromagnetic force F and the target pressure
difference TPD. After step S62, the interruption routine is
terminated and the controller 70 restarts engine control.
[0079] The flowchart in FIG. 9 shows irregular interruption routine
2 which is effective when the A/C switch 81 is ON. When the signal
output from the detection circuit 85 changes, the controller 70
determines that an interruption request has been made and
interrupts engine control to initiate irregular interruption
routine 2. When an ON signal is input in step S81, the controller
70 performs a regular interruption routine A, which is shown in
FIG. 10, in step S82. When a falling signal is input in S81, the
controller 70 performs regular interruption routine B, which is
shown in FIG. 11, in step S83. After step S82 or S83, the
interruption routine is terminated and the controller 70 restarts
engine control.
[0080] For example, the instant the temperature near the evaporator
33 drops and the monitored temperature falls below the lower
temperature limit T1 due to the discharge of the coolant from the
compressor based on the aforementioned initial value DtIni, the
controller 70 receives an OFF or falling signal from the detection
circuit 85. Then, the controller 70 regularly executes regular
interruption routine B, which is shown in FIG. 11, until the
routine is changed to routine A upon the next reception of an ON or
rising signal. Regular interruption routine B is carried out in
synchronism with the clock signal from the clock signal
generator.
[0081] When the controller 70 interrupts engine control and starts
routine B, the controller 70 reduces the present duty ratio Dt by a
unit amount .DELTA.D, which is shown in FIG. 11, in step S101. When
the duty ratio Dt decreases, the target pressure difference TPD
decreases. In other words, the displacement of the compressor is
reduced. Consequently, air-conditioning is controlled to reduce
cooling.
[0082] Subsequently, the controller 70 determines in step S102
whether or not a corrected value Dt-.DELTA.D, or the duty ratio Dt
reduced by the unit amount .DELTA.D, is smaller than a preset lower
limit value DtMin. When the determination is NO in step S102, the
controller 70 commands the drive circuit 72 to change the duty
ratio Dt in step S103. Then, the electromagnetic force F of the
solenoid section is weakened slightly, which reduces the target
pressure difference TPD of the control valve CV accordingly. Then,
the actuation rod 40 moves downward, which causes the return spring
66 to expand. The new position of the actuation rod 40 is
determined by satisfaction of equation 5. As a result, the degree
of opening of the control valve CV, or the degree of opening of the
supply passage 28, increases, thus increasing the crank pressure
Pc. This increases the difference between the crank pressure Pc and
the pressure in a cylinder bore 1a via an associated piston 20. As
a result, the inclination angle of a swash plate 12 decreases,
which reduces the displacement of the compressor. As the
displacement of the compressor decreases, the heat removing
performance in the evaporator 33 decreases, which reduces the
differential pressure between pressure monitor points P1 and
P2.
[0083] When the determination is YES in step S102, the controller
70 changes the duty ratio Dt to the lower limit value DtMin of, for
example, 0% in step S104. Then, the controller 70 instructs the
drive circuit 72 to perform duty control with the lower limit value
DtMin in step S103.
[0084] As regular interruption routine B is repeated, the duty
ratio Dt (i.e., the target pressure difference TPD) is reduced
decrementally as the time passes. A time chart in FIG. 13 shows a
time-dependent change in duty ratio Dt when regular interruption
routine B is repeated. The controller 70 reduces the duty ratio Dt
by the decremental amount .DELTA.D at a time in accordance with the
clock pulse signal until it next receives the ON signal after
reception of the OFF signal from the detection circuit 85.
Repeating this gradual reduction of the duty ratio Dt by the unit
amount .DELTA.D causes the duty ratio Dt to slowly decrease toward
the lower limit value DtMin (see the range between t3 and t4 of the
graph in FIG. 13). The controller 70 keeps the duty ratio Dt at the
lower limit value DtMin until it receives the ON signal from the
detection circuit 85 (see the range of the graph in FIG. 13
starting at t4).
[0085] When the reduction in the duty ratio Dt reduces the
displacement of the compressor and lowers the heat removing
performance in the evaporator 33, the temperature in the vehicle or
the monitored temperature gradually rises. When the monitored
temperature rises above the upper temperature limit T2, the
controller 70 receives the ON signal from the detection circuit 85.
Then, the controller 70 repeats regular interruption routine A
shown in FIG. 10 until it receives the OFF signal.
[0086] When the controller 70 interrupts engine control and starts
routine A, the controller 70 increases the present duty ratio Dt by
the unit amount .DELTA.D in step S91. When the duty ratio Dt
increases, the target pressure difference TPD increases. In other
words, the displacement of the compressor increases. Consequently,
the air-conditioning is controlled to increase the cooling
performance.
[0087] Subsequently, the controller 70 determines in step S92
whether or not a corrected value Dt+.DELTA.D of the duty ratio Dt
is greater than an upper limit value DtMax(x). The upper limit
value DtMax(x) is initially set at DtMax(100) in irregular
interruption routine 1, as shown in FIG. 7, before a predetermined
time has passed since the A/C switch 81 was switched ON. When the
predetermined time has passed, on the other hand, the upper limit
value DtMax(x) is changed to DtMax(70) in regular interruption
routine C, as shown in FIG. 8.
[0088] When the upper limit value DtMax(x) is set to DtMax(100) in
step S92, the controller 70 merely monitors whether the duty ratio
Dt has been computed to be greater than the actual control range (0
to 100%) of the drive signal output from the drive circuit 72. When
the controller 70 instructs the drive circuit 72 to set the duty
ratio Dt higher than DtMax(100), for example, the target pressure
difference TPD is set to DtMax(100). The following is one of the
reasons why the controller 70 does not allow the duty ratio Dt to
become beyond DtMax(100). Even if regular interruption routine B
becomes effective with the duty ratio Dt set above DtMax(100) and
the duty ratio Dt is gradually reduced, the target pressure
difference TPD is kept at the maximum value until the duty ratio Dt
becomes lower than DtMax(100).
[0089] The same is true of the case where the duty ratio Dt becomes
smaller than DtMin(0%). In step S102, therefore, the controller 70
determines whether or not the duty ratio Dt lies below the actual
control range (0-100%) of the drive circuit 72.
[0090] When the upper limit value DtMax(x) is set to DtMax(70) in
step S92, the controller 70 determines whether or not the target
pressure difference TPD calculated in step S91 is equal to or
higher than the upper limit value.
[0091] When the determination is NO in step S92, the controller 70
commands the drive circuit 72 to change the duty ratio Dt in step
S93. Then, the electromagnetic force F of the solenoid section
increases, thus increasing the target pressure difference TPD of
the control valve CV accordingly. As a result, the actuation rod 40
moves upward, which causes the return spring 66 to be compressed.
The actuation rod 40 is shifted such that the downward urging force
f2 of the return spring 66 is increased to offset the increase of
the upward electromagnetic force F. In other words, the actuation
rod 40 shifts to a position where the equation 5 is satisfied. As a
result, the degree of opening of the control valve CV, or the
degree of opening of the supply passage 28 decreases, thus lowering
the crank pressure Pc. This reduces the difference between the
crank pressure Pc and the pressure in the cylinder bore 1a. Then,
the inclination angle of the swash plate 12 increases, which
increases the displacement of the compressor. As the displacement
of the compressor increases, the heat removing performance in the
evaporator 33 is also increased, and the monitored temperature
falls. Therefore, the differential pressure between the pressure
monitor points P1 and P2 increases.
[0092] When the determination is YES in step S92, the controller 70
changes the duty ratio Dt to the upper limit value DtMax(x) in step
S94. When the upper limit value DtMax(x) is set to DtMax(100), the
controller 70 merely corrects the duty ratio Dt computed in step
S91 so that the duty ratio Dt falls within the actual control range
(0-100%) of the drive circuit 72 output from the drive circuit 72.
Accordingly, the substantial target pressure difference TPD
indicated by the duty ratio Dt is maintained. On the other hand,
when the upper limit value DtMax(x) is set to DtMax(70), the
controller 70 reduces the target pressure difference TPD calculated
to be above the upper limit value DtMax(70) in step S91 to the
value that matches with the upper limit value DtMax(70). In other
words, the target pressure difference TPD is changed to a new TPD
that corresponds to the upper limit value DtMax(70). As the flow
proceeds to step S93 from step S94, the controller 70 instructs the
drive circuit 72 to perform duty control with the upper limit value
DtMax(x).
[0093] As regular interruption routine A is repeated, the duty
ratio Dt and the target pressure difference TPD increase as the
time passes.
[0094] A time chart in FIG. 12 shows a time-dependent change in
duty ratio Dt when regular interruption routine A is repeated. The
controller 70 increases the duty ratio Dt by the unit amount
.DELTA.D incrementally in accordance with the clock pulse signal
from the clock signal generator until it receives the OFF signal
after reception of the ON signal from the detection circuit 85.
Repeatedly increasing the duty ratio Dt by the unit amount D causes
the duty ratio Dt to keep increasing slowly with the upper limit
value DtMax(x) as the maximum (see the portion between t1 and t2 of
the graph in FIG. 12 (when the upper limit value DtMax(x) is
DtMax(100)) or t1 to t2' (when the upper limit value DtMax(x) is
DtMax(70) indicated by the two-dot chain line)). The controller 70
keeps the duty ratio Dt at the upper limit value DtMax(x) until it
receives the OFF signal from the detection circuit 85 (see the
portion of the graph in FIG. 12 starting at t2 (when the upper
limit value DtMax(x) is DtMax(100)) or starting at t2 (when the
upper limit DtMax(x) is DtMax(70) indicated by the two-dot chain
line)).
[0095] When the increase in the duty ratio Dt increases the
displacement of the compressor and increases the heat removing
performance in the evaporator 33, the temperature in the vehicle,
or the monitored temperature, gradually falls. When the monitored
temperature falls below the lower temperature limit T1, the
controller 70 repeats regular interruption routine B until it
receives the ON signal.
[0096] In other words, the controller 70 continues the process of
gradually increasing or decreasing the duty ratio Dt, which
represents the target pressure difference TPD of the control valve
CV. When receiving the detection circuit signal from the detection
circuit 85, the controller 70 changes the target pressure
difference TPD. Increasing and decreasing the target pressure
difference TPD (duty ratio Dt) are alternately repeated in this
way. Without a sudden change in cooling load, the duty ratio Dt
shows a time-dependent change as indicated by a solid line 131 in a
time chart in FIG. 14. When the monitored temperature changes
between the threshold temperatures T1 and T2, the rising signal and
the falling signal from the detected circuit 85 switches. According
to the switching the duty ratio Dt repeats the alternate increase
and decrease while nearly keeping a constant fluctuation with
respect to the center value, DtMid(t). That is, the ON-OFF control
by the controller 70 regulates the duty ratio Dt close to the
center value DtMid(t). In this embodiment, DtMid(t) is a variable
that varies with the passing of time but is substantially constant
as indicated by a one-dot chain line 132 in FIG. 14, for
example.
[0097] Even if the thermal load of the evaporator 33 varies, the
duty ratio Dt, or the target pressure difference TPD, is optimized
for controlling the circulation amount of the coolant. This allows
the temperature near the evaporator 33 to be kept at the optimal
temperature.
[0098] This embodiment has the following advantages.
[0099] (1) The suction pressure Ps, which is influenced by the
level of the thermal load of the evaporator 33, is not directly
used as an index for controlling the degree of opening of the
displacement control valve CV. The differential pressure
.DELTA.P(t)-PdH-PdL between two pressure monitor points P1 and P2
in the cooling circuit is the direct control target in the feedback
control of the displacement of the compressor. This enables the
displacement to be quickly reduced in the displacement regulation
control according to the value of the current supplied
independently of the thermal load on the evaporator 33.
[0100] (2) The operational efficiency of the compressor decreases
lower as the piston speed increases due to an increase in friction.
The piston speed is associated with the rotational speed of the
drive shaft 6, which has a specific relation with the engine speed
NE of the vehicle's engine E, and the displacement of the
compressor (which determines the stroke of the piston 20). The
compressor cannot change the engine speed NE. To use the compressor
efficiently or to improve the operational efficiency of the engine
E, therefore, the displacement should not be maximized when the
engine speed NE is high. To achieve this, the differential pressure
.DELTA.P(t)=PdH-PdL between the two points when the displacement of
the compressor is maximum and when the engine speed NE is low
should be the maximum value of the target pressure difference TPD
used when the duty ratio Dt is DtMax(100). Through this control
procedure, when the engine speed NE becomes high, the differential
pressure .DELTA.P(t) always exceeds the maximum value of the target
pressure difference TPD when the displacement is maximum, and the
compressor automatically reduces the displacement from the
maximum.
[0101] At the initial stage of cooling, an air-conditioning system
for a vehicle must produce maximum cooling performance regardless
of the engine speed NE. For vehicle air-conditioning systems,
therefore, it is desirable to design the control valve in
consideration of the initial stage of cooling, rather than high
efficiency and light-load operation. That is, the control valve is
designed in such a way that the differential pressure
.DELTA.P(t)=PdH-PdL between the two points when the displacement of
the compressor is maximum and when the engine speed NE is low is
the maximum value of the target pressure difference TPD. With such
a design, even when the displacement is maximum, no matter how high
the engine speed NE is, the differential pressure
.DELTA.P(t)=PdH-PdL does not rise above the maximum value of the
target pressure difference TPD. When the duty ratio Dt is set to
DtMax(100), therefore, the displacement of the compressor is always
maximum. This allows the vehicle air-conditioning system to produce
the maximum cooling performance at that point, regardless of the
engine speed NE, so that the requirement for rapid cooling at the
initial stage of cooling is met.
[0102] Because of the aforementioned design of the control valve,
which prioritizes the control at the initial stage of cooling, if
regular interruption routine C in FIG. 8 were not provided, the
following problem would occur. Suppose that a predetermined time
has elapsed since the switching of the A/C switch 81 to the ON
state from the OFF state and the monitored temperature has dropped
to a predetermined temperature. In this state, the requirement for
rapid cooling at the initial stage of cooling is mostly met and the
maximum cooling performance need not be performed thereafter to
keep the monitored temperature at the threshold temperature. More
specifically, after the requirement for rapid cooling is met, even
if the cooling load is heavy, it is possible to provide an adequate
amount of coolant without maximizing the displacement of the
compressor as long as the engine speed NE is equal to or higher
than the vehicle speed V of, for example, 40 km/h at the top gear
ratio, i.e., even when the engine speed NE is low. This permits the
monitored temperature to be maintained at the threshold temperature
without difficulty.
[0103] When a control that lacks regular interruption routine C in
FIG. 8 is executed, however, use of the duty ratio Dt at the upper
limit value DtMax(100) is always permitted. Even after the
requirement for rapid cooling at the initial cooling stage is met,
therefore, the duty ratio Dt may be set to the upper limit value
DtMax(100) at the end of routine A. When the engine speed NE is
high, the aforementioned characteristics of the control valve cause
the displacement of the compressor to be maximized so that the
air-conditioning system maintains the maximum cooling performance.
If this state occurs after the requirement for rapid cooling at the
initial cooling stage is met, the compressor functions
unnecessarily. This results in low efficiency.
[0104] According to this embodiment, however, when a predetermined
time has elapsed from when the OFF-to-ON switching of the A/C
switch 81, the controller 70 determines that the monitored
temperature has dropped to a certain level where the requirement
for rapid cooling at the initial cooling stage is met, and changes
the duty ratio Dt to the upper limit value DtMax(70). After the
predetermined time has elapsed since the ON action of the A/C
switch 81, the target pressure difference TPD does not rise above
the upper limit value determined by the reduced duty ratio. Should
the target pressure difference TPD be set to the upper limit value,
the differential pressure .DELTA.P(t) between the two monitor
points always rises above the upper limit value when the engine
speed NE becomes high. Accordingly, the displacement of the
compressor automatically falls from the maximum value. Therefore,
the compressor does not operate as low efficiency under a high-load
unnecessarily, and the operational efficiency of the engine E is
improved, which improves fuel consumption. It is also possible to
use the compressor over a longer period of time. When the engine
speed NE is high or when the engine E is in a high-load state, the
displacement of the compressor (load torque) is not maximized.
Therefore, the drive load of the engine E is reduced, thus
improving acceleration and vehicle performance at high speeds.
Since, as a consequence, the amount of heat from the engine E is
reduced, the engine cooling system (particularly, the heat
exchanger) can be smaller.
[0105] (3) The temperature near the evaporator 33 is kept at the
optimal temperature for cooling by a simple control procedure to
increase and decrease the duty ratio Dt based on the rising signal
and falling signal input from the detection circuit 85. That is,
the burden of computation on the control unit is reduced by the use
of a control sequence that is simple enough to be handled by an
interruption routine. This allows the controller 70, which also
controls the engine E, to maintain the temperature of the passenger
compartment. This eliminates the need for an expensive control unit
exclusively for the air-conditioner. This lowers the manufacturing
cost of the compressor.
[0106] (4) The lower and upper temperature limits T1 and T2 are set
as threshold temperatures to provide a hysteresis, which makes the
temperature at which the rising signal is output from the detection
circuit 85 different from the temperature at which the falling
signal is output. This feature avoids hunting, which is apt to
result when a single threshold temperature is used, thus
stabilizing the displacement control. Hunting in the detection
circuit 85 results in frequent generation of the detection circuit
signal that indicates the reversing of the level relationship
between the monitored temperature and the single threshold
temperature.
[0107] (5) Since the movable wall 54 and the actuation rod 40
respond to the pressures PdH and PdL at the two pressure monitor
points P1 and P2, the force based on the differential pressure
.DELTA.P(t)=PdH-PdL is applied to the valve body 43. This
embodiment does not therefore require a complicated structure
(pressure sensors or the like) that electrically detects the two
pressures PdH and PdL at the two pressure monitor points P1 and P2,
for example, or a program for electrical control of the coil 67
(drive circuit 72).
[0108] (6) The compressor is a variable displacement type swash
plate compressor designed to change the stroke of each piston 20 by
controlling the crank pressure Pc. The control apparatus of the
embodiment is most suitable for displacement control of such a
variable displacement type swash plate compressor.
[0109] The present invention includes the following
embodiments.
[0110] Only a single threshold temperature may be set as the
threshold temperature instead of setting an upper temperature limit
and a lower temperature limit, which are different.
[0111] The first pressure monitor point P1 may be located in a
suction pressure area between the evaporator 33 and a suction
chamber 21, and the second pressure monitor point P2 may be
provided at the downstream of the first pressure monitor point
P1.
[0112] The first pressure monitor point P1 may be located in the
discharge pressure area between the discharge chamber 22 and the
condenser 31, and the second pressure monitor point P2 may be
located in the suction pressure area between the evaporator 33 and
the suction chamber 21.
[0113] The first pressure monitor point P1 may be located in the
discharge pressure area between the discharge chamber 22 and the
condenser 31, and the second pressure monitor point P2 may be
located in the crank chamber 5. Alternatively, the first pressure
monitor point P1 may be in the crank chamber 5, and the second
pressure monitor point P2 may be in the suction pressure area
between the evaporator 33 and the suction chamber 21. The locations
of the first and second pressure monitor points P1 and P2 are not
limited to the coolant passage that is the main passage of the
cooling circuit and are not limited to the evaporator 33, the
suction chamber 21, the cylinder bore 1a, the discharge chamber 22
and the condenser 31. That is, the location of each pressure
monitor point P1 or P2 is not limited to the high-pressure area or
the low-pressure area in the coolant passage. For example, the
pressure monitor points P1 and P2 may be provided in the coolant
passage for displacement control, which is the sub circuit of the
cooling circuit, i.e., the two points P1 and P2 may be in the crank
chamber 5 or an intermediate-pressure area among the air intake
passage 28, the crank chamber 5 and a bleeder passage 27.
[0114] In the latter case, when the displacement of the compressor
increases, the differential pressure .DELTA.P(t)=Pc-Ps falls. If
the elapsed time is equal to or greater than a predetermined time,
therefore, the differential pressure .DELTA.P(t) between the two
pressure monitor points is set to the lower limit. Then, the
target-differential-pressure determiner compares the target
pressure difference computed by the target-differential-pressure
calculator with the lower limit, and determines the target pressure
difference as a new target pressure difference when the target
pressure difference is equal to or higher than the lower limit or
determines the lower limit as a new target pressure difference when
the target pressure difference is lower than the lower limit.
[0115] The control valve may be electrically driven, and the
pressures PdH and PdL at the two pressure monitor points P1 and P2
may be detected by associated pressure sensors.
[0116] The control valve may be a so-called outlet control valve,
which regulates the crank pressure Pc by adjusting the degree of
opening of the bleed passage 27.
[0117] The control valve may be a three-way valve that regulates
the crank pressure Pc by adjusting the opening size of both the air
intake passage 28 and the bleed passage 27.
[0118] The power transmission mechanism PT may be equipped with a
clutch mechanism such as an electromagnetic clutch.
[0119] The present invention may be embodied into a control
apparatus for a wobble type variable displacement type
compressor.
[0120] 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.
[0121] 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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