Adaptive control for a refrigeration system using pulse width modulated duty cycle scroll compressor

BACKGROUND AND SUMMARY OF THE INVENTION

The present invention relates generally to refrigeration systems, compressor control systems and refrigerant regulating valve control systems. More particularly, the invention relates to a refrigeration system employing a pulse width modulated compressor or evaporator stepper regulator controlled by a variable duty cycle signal derived from a load sensor. Preferably an adaptive controller generates the variable duty cycle signal. The compressor has two mechanical elements separated by a seal, and these mechanical elements are cyclically movable relative to one another to develop fluid pressure. The compressor includes a mechanism to selectively break the seal in response to the control signal, thereby modulating the capacity of the system.

The refrigeration system can be deployed as a distributed system in refrigeration cases and the like. The preferred arrangement allows the compressor and condenser subsystems to be disposed in or mounted on the refrigeration case, thereby greatly reducing the length of refrigerant conduit and refrigerant required.

Conventionally, refrigeration systems for supermarket refrigeration cases have employed air-cooled or water-cooled condensers fed by a rack of compressors. The compressors are coupled in parallel so that they may be switched on and off in stages to adjust the system cooling capacity to the demands of the load. Commonly, the condensers are located outside, on the roof, or in a machine room adjacent the shopping area where the refrigeration cases are located.

Within each refrigeration case is an evaporator fed by lines from the condensers through which the expanded refrigerant circulates to cool the case. Conventionally, a closed-loop control system regulates refrigerant flow through the evaporator to maintain the desired case temperature. Proportional-integral-derivative (PID) closed loop control systems are popular for this purpose, with temperature sensors and/or pressure sensors providing the sensed condition inputs.

It is common practice within supermarkets to use separate systems to supply different individual cooling temperature ranges: low temperature (for frozen foods, ice cream, nominally −25 F.); medium (for meat, dairy products, nominally +20 F.); high (for floral, produce, nominally +35 to +40 F.). The separate low, medium and high temperature systems are each optimized to their respective temperature ranges. Normally, each will employ its own rack of compressors and its own set of refrigerant conduits to and from the compressors and condensers.

The conventional arrangement, described above, is very costly to construct and maintain. Much of the cost is associated with the long refrigerant conduit runs. Not only are long conduit runs expensive in terms of hardware and installation costs, but the quantity of refrigerant required to fill the conduits is also a significant factor. The longer the conduit run, the more refrigerant required. Adding to the cost are environmental factors. Eventually fittings leak, allowing the refrigerant to escape to atmosphere. Invariably, long conduit runs involve more pipefitting joints that may potentially leak. When a leak does occur, the longer the conduit run, the more refrigerant lost.

There is considerable interest today in environmentally friendly refrigeration systems. Shortening the conduit run is seen as one way to achieve a more environmentally friendly system. To achieve this, new condenser/compressor configurations and new control systems will need to be engineered.

Re-engineering condenser/compressor configurations for more environmentally friendly systems is not a simple task, because system efficiency should not be sacrificed. Generally, the conventional roof-mounted condenser system, supplied by condensers, benefits from economies of scale and is quite efficient. These systems serve as the benchmark against which more environmentally friendly systems of the future will need to be measured.

To appreciate why re-engineering an environmentally yet efficient system has proven so difficult, consider these thermodynamic issues. The typical refrigeration case operates in a very unpredictable environment. From a design standpoint, the thermal mass being cooled is rarely constant. Within the supermarket environment, the temperature and humidity may vary widely at different times of day and over different seasons throughout the year. The product load (items in the refrigeration case) can also change unpredictably. Customers removing product and store clerks replenishing product rarely synchronize. Outside the supermarket environment, the outdoor air temperature and humidity may also vary quite widely between day and night and/or between summer and winter. The capacity of the system must be designed for the harshest conditions. (when the condenser environment is the hottest). Thus systems may experience excess capacity in less harsh conditions, such as in the cool evenings or during the winter.

Periodic defrosting also introduces thermal fluctuations into the system. Unlike thermal fluctuations due to environmental conditions, the thermal fluctuations induced by the defrost cycle are cause by the control system itself and not by the surrounding environment.

In a similar fashion, the control system for handling multiple refrigeration cases can induce thermal fluctuations that are quite difficult to predict. If all cases within a multi-case system are suddenly turned on at once—to meet their respective cooling demands—the cooling capacity must rapidly be ramped up to maximum. Likewise, if all cases are suddenly switched off, the cooling capacity should be ramped down accordingly. However, given that individual refrigeration cases may operate independently of one another, the instantaneous demand for cooling capacity will tend to vary widely and unpredictably.

These are all problems that have made the engineering of environmentally friendly systems more difficult. Adding to these difficulties are user engineering/ergonomic problems. The present day PID controller can be difficult to adapt to distributed refrigeration systems. Experienced controls engineers know that a well-tuned PID controller can involve a degree of artistry in selecting the proper control constants used in the PID algorithm. In a large refrigeration system of the conventional architecture (non-distributed) the size of the system justifies having a controls engineer visit the site (perhaps repeatedly) to fine tune the control constant parameters.

This may not be practical for distributed systems in which the components are individually of a much smaller scale and far more numerous. By way of comparison, a conventional system might employ one controller for an entire multi-case, store-wide system. A distributed system for the same store might involve a controller for each case or adjacent group of cases within the store. Distributed systems need to be designed to minimize end user involvement. It would therefore be desirable if the controller were able to auto configure. Currently control systems lack this capability.

The present invention provides a distributed refrigeration system in which the condenser is disposed on the refrigeration case and serviced by a special pulse width modulated compressor that may be also disposed within the case. If desired, the condenser and compressor can be coupled to service a group of adjacent refrigeration cases, each case having its own evaporator. The pulse width modulated compressor employs two mechanical elements, such as scroll members, that move rotationally relative to one another to develop fluid pressure for pumping the refrigerant. The compressor includes a mechanism that will selectively break the seal between the two mechanical elements, thereby altering the fluid pressure developed by the compressor while allowing the mechanical elements to maintain substantially constant relative movement with one another. The compressor can be pulse width modulated by making and breaking the fluid seal without the need to start and stop the electric motor driving the mechanical elements.

The pulse width modulated compressor is driven by a control system that supplies a variable duty cycle control signal based on measured system load. The controller may also regulate the frequency (or cycle time) of the control signal to minimize pressure fluctuations in the refrigerant system. The on time is thus equal to the duty cycle multiplied by the cycle time, where the cycle time is the inverse of the frequency.

The refrigeration system of the invention has a number of advantages. Because the instantaneous capacity of the system is easily regulated by variable duty cycle control, an oversized compressor can be used to achieve faster temperature pull down at startup and after defrost, without causing short cycling as conventional compressor systems would. Another benefit of variable duty cycle control is that the system can respond quickly to sudden changes in condenser temperature or case temperature set point. The controller adjusts capacity in response to disturbances without producing unstable oscillations and without significant overshoot. Also, the ability to match instantaneous capacity to the demand allows the system to operate at higher evaporator temperatures. (Deep drops in temperature experienced by conventional systems at overcapacity are avoided.)

Operating at higher evaporator temperatures reduces the defrost energy required because the system develops frost more slowly at higher temperatures. Also, the time between defrosts can be lengthened by a percentage proportional to the accumulated runtime as dictated by the actual variable duty cycle control signal. For example, a sixty percent duty cycle would increase a standard three-hour time between defrosts to five hours (3/0.60=5).

The pulse width modulated operation of the system yields improved oil return. The refrigerant flow pulsates between high capacity and low capacity (e.g. 100% and 0%), creating more turbulence which breaks down the oil boundary layer in the heat exchangers.

Another benefit of the variable duty cycle control system is its ability to operate with a variety of expansion devices, including the simple orifice, the thermal expansion valve (TXV) and the electronic expansion valve. A signal derived from the expansion device controller can be fed to the compressor controller of the invention. This signal allows the variable duty cycle control signal and/or its frequency to be adjusted to match the instantaneous operating conditions of the expansion device. A similar approach may be used to operate variable speed fans in air cooled condenser systems. In such case the controller of the invention may provide a signal to control fan speed based on the current operating duty cycle of the compressor.

Yet another benefit of the invention is its ability to detect when the system is low on refrigerant charge, an important environmental concern. Low refrigerant charge can indicate the presence of leaks in the system. Low charge may be detected by observing the change in error between actual temperature and set point temperature as the system duty cycle is modulated. The control system may be configured to detect when the modulation in duty cycle does not have the desired effect on temperature maintenance. This can be due to a loss of refrigerant charge, a stuck thermal expansion valve or other malfunctions.

For a more complete understanding of the invention, its objects and advantages, refer to the following specification and to the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1

is a system block diagram of a prior art refrigeration system-configuration;

FIG. 2

is a block diagram of a refrigeration system in accordance with the present invention;

FIG. 3

is a cross-sectional view of an embodiment of the pulse width modulated compressor, shown in the loaded state;

FIG. 4

is a cross-sectional view of the compressor of

FIG. 3

, shown in the unloaded state;

FIG. 5

is another embodiment of a refrigeration or cooling system in accordance with the present invention;

FIG. 6

is a block diagram of the controller;

FIG. 7

is a block diagram showing how the controller may be used to modulate an evaporator stepper regulator;

FIG. 8

is a block diagram of the signal conditioning module of the controller of

FIG. 6

;

FIG. 9

is a block diagram of the control module of the controller of

FIG. 6

;

FIG. 10

is a state diagram depicting the operating states of the controller;

FIG. 11

is a flowchart diagram illustrating the presently preferred PI control algorithm;

FIG. 12

is a waveform diagram illustrating the variable duty cycle signal produced by the controller and illustrating the operation at a constant frequency;

FIG. 13

is a waveform diagram of the variable duty cycle signal, illustrating variable frequency operation;

FIG. 14

is a series of graphs comparing temperature and pressure dynamics of system employing the invention with a system of conventional design;

FIG. 15

is a block diagram illustrating the adaptive tuning module of the invention;

FIG. 16

a

is a flowchart diagram illustrating the presently preferred operation of the adaptive tuning module, specifically with respect to the decision whether to start tuning;

FIG. 16

b

is a flowchart diagram illustrating the presently preferred process performed by the adaptive tuning module in the integration mode;

FIG. 16

c

is a flowchart diagram illustrating the operation of the adaptive tuning module in the calculation mode;

FIG. 17

is a state diagram illustrating the operative states of the adaptive tuning module;

FIG. 18

is a block diagram illustrating the fuzzy logic block of the adaptive tuning loop;

FIG. 19

is a membership function diagram for the fuzzy logic block of

FIG. 18

;

FIG. 20

is a truth table relating to the membership function of

FIG. 19

as used by the fuzzy logic block of

FIG. 18

;

FIG. 21

is an output membership function diagram for the fuzzy logic block of

FIG. 18

; and

FIG. 22

is a schematic illustrating exemplary sensor locations for control-related and diagnostic-related functions of the present invention.

DESCRIPTION OF THE PREFERRED EMBODIMENTS

FIG. 1

illustrates how a conventional supermarket refrigeration system is configured. As previously discussed, it is conventional practice to place the compressors

30

and the condenser

32

in a location remote from the refrigeration cases

34

. In this illustration, the compressors

30

are configured in a parallel bank located on the roof

36

of the building. The bank of compressors supply a large condenser

32

, which may be air cooled or water cooled. The condenser supplies liquid refrigerant to a receiver

38

. The receiver

38

, in turn, supplies the individual refrigeration cases, which are connected in parallel, as illustrated. In most implementations a liquid line solenoid valve

40

is used to regulate flow to the associated evaporator

42

. The refrigerant is supplied to the evaporator through a suitable expansion device such as expansion valve

44

. The expansion valve

44

provides a restricted orifice that causes the liquid refrigerant to atomize into liquid droplets that are introduced into the inlet side of the evaporator

42

. The evaporator

42

, located within the refrigeration case

34

, extracts heat from the case and its contents by vaporization of the liquid refrigerant droplets into a gas. The compressors

30

extract this gas by suction and compress it back into the liquid state. The liquid refrigerant is then cooled in the condenser

32

and returned to the receiver

38

, whereupon the cycle continues.

To match cooling capacity to the load, the compressors

30

may be switched on and off individually or in groups, as required. In a typical supermarket arrangement there may be several independent systems, each configured as shown in

FIG. 1

, to handle different operating temperature ranges. Note that the liquid line

46

and the suction line

48

may each need to be quite lengthy (e.g., up to 150 feet) to span the distance from refrigeration case to roof.

FIG. 2

shows a refrigeration case

34

configured according to the principles of the present invention. The condenser

32

and compressor

30

are both disposed within case

34

or attached thereto. Evaporator

42

and the associated expansion valve

44

are likewise disposed within case

34

. The condenser

32

is provided with a heat removal mechanism

50

by which heat is transferred to ambient. The heat removal mechanism can be a water jacket connected to suitable plumbing for carrying waste heat to a water cooling tower located on the building roof. Alternatively, the heat removal mechanism can be a forced-air cooling system or a passive convection-air cooling system.

The refrigeration system of the invention employs a compressor controller

52

that supplies a pulse width modulated control signal on line

54

to a solenoid valve

56

on compressor

30

. The compressor controller adjusts the pulse width of the control signal using an algorithm described below. A suitable load sensor such as temperature sensor

58

supplies the input signal used by the controller to determine pulse width.

FIGS. 3 and 4

show the details of compressor

30

.

FIG. 3

shows the compressor in its loaded state and

FIG. 4

shows the compressor in its unloaded state. The solenoid valve

56

switches the compressor between these two states while the compressor motor remains energized. One important advantage of this configuration is that the compressor can be pulse width modulated very rapidly between the loaded and unloaded states without interrupting power to the compressor motor. This pulse width modulated cycling exerts less wear on the compressor, because the motor is not subjected to sudden changes in angular momentum.

Referring to

FIGS. 3 and 4

, there is shown an exemplary compressor

30

. Compressor

30

may be used within a hermetic scroll compressor such as generally of the type described in assignee's U.S. Pat. No. 5,102,316.

The exemplary compressor

30

includes an outer shell

61

and an orbiting scroll member

64

supported on upper bearing housing

63

and drivingly connect to crankshaft

62

via crank pin

65

and drive bushing

60

. A second non-orbiting scroll member

67

is positioned in meshing engagement with scroll member

64

and axially movably secured to upper bearing housing

63

. A partition plate

69

is provided adjacent the upper end of shell

61

and serves to define a discharge chamber

70

at the upper end thereof.

In operation, as orbiting scroll member

64

orbits with respect to scroll member

67

, suction gas is drawn into shell

61

via suction inlet

71

and thence into compressor

30

through inlet

72

provided in non-orbiting scroll member

67

. The intermeshing wraps provided on scroll members

64

and

67

define moving fluid pockets which progressively decrease in size and move radially inwardly as a result of the orbiting motion of scroll member

64

thus compressing the suction gas entering via inlet

72

. The compressed gas is then discharged into discharge chamber

70

via discharge port

73

provided in scroll member

67

and passage

74

.

In order to unload compressor

30

, solenoid valve

56

will be actuated in response to a signal from control module

87

to interrupt fluid communication to increase the pressure within chamber

77

to that of the discharge gas. The biasing force resulting from this discharge pressure will overcome the sealing biasing force thereby causing scroll member

67

to move axially upwardly away from orbiting scroll member

64

. This axial movement will result in the creation of a leakage path between the respective wrap tips and end plates of scroll members

64

and

67

thereby substantially eliminating continued compression of the suction gas.

A flexible fluid line

91

extends from the outer end of passage

90

to a fitting

92

extending through shell

61

with a second line

93

connecting fitting

92

to solenoid valve

56

. Solenoid valve

56

has fluid lines

82

and

84

connected to suction line

83

and discharge line

85

and is controlled by control module

87

in response to conditions sensed by sensor

88

to effect movement of non-orbiting scroll member

67

between the positions shown in

FIGS. 3 and 4

.

When compression of the suction gas is to be resumed, solenoid valve

56

will be actuated so as to move scroll member

67

into sealing engagement with scroll member

64

.

The refrigeration case embodiment of

FIG. 2

may be packaged as a self-contained unit. While that may be a desirable configuration for many applications, the invention is not restricted to stand alone, self-contained refrigeration case configurations. Rather, the invention lends itself to a variety of different distributed refrigeration systems.

FIG. 5

shows an example of such a distributed system.

Referring to

FIG. 5

, a single compressor

30

and condenser

32

can service several distributed refrigeration cases or several distributed cooling units in a heating and cooling (HVAC) system. In

FIG. 5

the refrigeration cases or cooling system housings are shown as dashed boxes, designated

34

a,

34

b,

and

34

c.

Conveniently, the compressor

30

and condenser

32

may be disposed within or attached to one of the refrigeration cases or housings, such as refrigeration case or housing

34

a.

Each refrigeration case or housing has its own evaporator and associated expansion valve as illustrated at

42

(

a, b, c

) and

44

(

a, b, c

). In addition, each refrigeration case or housing may have its own temperature sensor

58

(

a, b, c

) supplying input information to the compressor controller

52

. Finally, a pressure sensor

60

monitors the pressure of the suction line

48

and supplies this information to compressor controller

52

. The compressor controller supplies a variable duty cycle signal to the solenoid valve

56

as previously described.

The multiple case or multiple cooling unit embodiment of

FIG. 5

shows how a single compressor can be pulse width modulated by compressor controller

52

to supply the instantaneous demand for cooling. The temperature sensors

58

(

a, b, c

) collectively provide an indication of the load on the system, as does pressure sensor

60

. The controller adjusts the pulse width of the control signal to modulate the compressor between its high capacity and low capacity states (100%, 0%) to meet the instantaneous demand for refrigerant.

As an alternate control technique, one or more of the suction lines exiting the evaporator can be equipped with an electrically controlled valve, such as an evaporator pressure regulator valve

45

c.

Valve

45

c

is coupled to controller

52

, as illustrated. It may be supplied with a suitable control signal, depending on the type of the valve. A stepper motor valve may be used for this purpose, in which case controller

30

would supply a suitable signal to increment or decrement the setting of the stepper motor to thereby adjust the orifice size of the valve. Alternatively, a pulse width modulated valve could be used, in which case it may be controlled with the same variable duty cycle signal as supplied to the compressor

30

.

Controller

52

is not limited to solely compressor control applications. The variable duty cycle control signal can also be used to control other types of refrigerant flow and pressure control devices, such as refrigerant regulating valves.

FIG. 7

shows such an application, where the output of controller

52

supplies control signals to evaporator stepper regulator

43

. This device is a fluid pressure regulator that is adjusted by stepper motor

45

. The evaporator stepper regulator (ESR) valve

43

adjusts the suction pressure to thereby adjust the capacity of the system.

A block diagram of the presently preferred compressor controller is illustrated in

FIG. 6. A

description of the various signals and data values shown in this and successive figures is summarized in Table I below.

TABLE I

Default

No.

Variable Name

Value

Description

1

Signal Conditioning:

Sensor Alarm

False

Indicates Sensor Reading is not

within expected range

Sensor Mode

Min

User configuration to indicate if

Min/Max/Avg is performed for

all temperature Sensors

Sampling Time (Ts)

0.5 sec

Rate at which Signal condition-

ing block is executed

Control Type

T/P Type

if controlled by only Temp. or

both Temp. & Pressure

2

Control Block:

Sensor Alarm

False

Same as before

System Alarm

False

Generated by Adaptive Block

indicative some system problem

SSL

0

Steady state loading %

Defrost Status

False

Whether system is in defrost

Pull_Down_Time

0

Time taken to pull down after

defrost

Gain (K)

7

Gain used in PI algorithm

Integral Time (Ti)

100

used in PI

Control Time (Tc)

10 Sec

used in PI

Control Set Pt. (St)

0 F.

used in PI

Operating State

1

What state the machine is

operating at

3

Defrost Control

Defrost Status

False

If defrost status of the case

Defrost Type

External

If the defrost is from external

timer or Internal clock of

controller

Defrost Interval

8 hrs

Time between defrost

Defrost Duration

1 hr

Defrost Duration

Defrost Termination

50 F.

Termination temperature for

Temp.

defrost

At the heart of the controller is control block module

102

. This module is responsible for supplying the variable duty cycle control signal on lead

104

. Module

102

also supplies the compressor ON/OFF signal on lead

106

and an operating state command signal on lead

108

. The compressor ON/OFF signal drives the contactors that supply operating current to the compressor motor. The operating state signal indicates what state the state machine (

FIG. 10

) is in currently.

The control block module receives inputs from several sources, including temperature and pressure readings from the temperature and pressure sensors previously described. These temperature readings are passed through signal conditioning module

110

, the details of which are shown in the pseudocode Appendix. The control block module also receives a defrost status signal from defrost control module

112

. Defrost control module

112

contains logic to determine when defrost is performed. The present embodiment allows defrost to be controlled either by an external logic signal (supplied through lead

114

) or by an internal logic signal generated by the defrost control module itself. The choice of whether to use external or internal defrost control logic is user selectable through user input.

116

. The internal defrost control uses user-supplied parameters supplied through user input

118

.

The preferred compressor controller in one form is autoconfigurable. The controller includes an optional adaptive tuning module

120

that automatically adjusts the control algorithm parameters (the proportional constant K) based on operating conditions of the system. The adaptive tuning module senses the percent loading (on lead

104

) and the operating state (on lead

108

) as well as the measured temperature after signal conditioning (on lead

122

). Module

120

supplies the adaptive tuning parameters to control block

102

, as illustrated. The current embodiment supplies proportional constant K on lead

124

and SSL parameter on lead

126

, indicative of steady-state loading percent. A system alarm signal on lead

126

alerts the control block module when the system is not responding as expected to changes in the adaptively tuned parameters. The alarm thus signals when there may be a system malfunction or loss of refrigerant charge. The alarm can trigger more sophisticated diagnostic routines, if desired. The compressor controller provides a number of user interface points through which user-supplied settings are input. The defrost type (internal/external) input

116

and the internal defrost parameters on input

118

have already been discussed. A user input

128

allows the user to specify the temperature set point to the adaptive tuning module

120

. The same information is supplied on user input

130

to the control block module

102

. The user can also interact directly with the control block module in a number of ways. User input

132

allows the user to switch the compressor on or off during defrost mode. User input

134

allows the user to specify the initial controller parameters, including the initial proportional constant K. The proportional constant K may thereafter be modified by the adaptive tuning module

120

. User input

136

allows the user to specify the pressure differential (dP) that the system uses as a set point.

In addition to these user inputs, several user inputs are provided for interacting with the signal conditioning module

110

. User input

138

selects the sensor mode of operation for the signal conditioning module. This will be described in more detail below. User input

140

allows the user to specify the sampling time used by the signal conditioning module. User input

142

allows the user to specify whether the controller shall be operated using temperature sensors only (T) or temperature and pressure sensors (T/P).

Referring now to

FIG. 8

, the signal conditioning module is shown in detail. The inputs (temperature and/or pressure sensors) are shown diagrammatically at

144

. These inputs are processed through analog to digital convertor

146

and then supplied to the control type selector

148

. Temperature readings from the temperature and/or pressure sensors are taken sequentially and supplied serially through the analog to digital convertor. The control type selector codes or stores the data so that pressure and temperature values are properly interpreted.

Digital filtering is then applied to the signal at

150

to remove spurious fluctuations and noise. Next, the data are checked in module

152

to ensure that all readings are within expected sensor range limits. This may be done by converting the digital count data to the corresponding temperature or pressure values and checking these values against the pre-stored sensor range limits. If the readings are not within sensor range an alarm signal is generated for output on output

154

.

Next a data manipulation operation is performed at

156

to supply the temperature and/or pressure data in the form selected by the sensor mode user input

138

. The current embodiment will selectively average the data or determine the minimum or maximum of the data (Min/Max/Avg). The Min/Max/Avg mode can be used to calculate the swing in pressure differential, or a conditioned temperature value. The average mode can be used to supply a conditioned temperature value. These are shown as outputs

158

and

160

, respectively.

FIG. 9

shows the control block module in greater detail. The conditioned temperature or pressure signal is fed to calculation module

162

that calculates the error between the actual temperature or pressure and the set point temperature or pressure. Module

162

also calculates the rate of change in those values.

The control block module is designed to update the operating state of the system on a periodic basis (every Tc seconds, nominally once every second). The Find Operating State module

164

performs this update function. The state diagram of

FIG. 10

provides the details on how this is performed. Essentially, the operating state advances, from state to state, based on whether there is a sensor alarm (SA) present, whether there is a defrost status signal (DS) present and what the calculated error value is. The Find Operating State module

164

supplies the operating state parameter and the Pull Down Time parameter to the decision logic module

166

.

Referring to

FIG. 10

, the Find Operating State module

164

advances from state to state as follows. Beginning in the initial state

168

the module advances to the normal operating state

170

after initialization. It remains in that state until certain conditions are met.

FIG. 10

shows by label arrows what conditions are required to cycle from the normal operating state

170

to the defrost state

172

; to the pull down state

174

; to the sensor alarm pull down state

176

; to the sensor alarm operating state

178

and to the sensor alarm defrost state

180

.

The decision logic module

166

(

FIG. 9

) determines the duty cycle of the variable duty cycle signal. This is output on lead

182

, designated % Loading. The decision logic module also generates the compressor ON/OFF signal on lead

184

. The actual decision logic will be described below in connection with FIG.

11

. The decision logic module is form of proportional integral (PI) control that is based on an adaptively calculated cycle time T

cyc

. This cycle time is calculated by the calculation module

186

based on a calculated error value generated by module

188

. Referring back to

FIG. 6

, the conditioned pressure differential signal on lead

122

(Cond dP) is supplied to the Calculate Error module

188

(

FIG. 9

) along with the pressure differential set point value as supplied through user input

136

(FIG.

6

). The difference between actual and set point pressure differentials is calculated by module

188

and fed to the calculation module

186

. The adaptive cycle time T

cyc

is a function of the pressure differential error and the operating state as determined by the find operating state module

164

according to the following calculation:

T

cyc(new)

=T

cyc(old)

+K

c

* Error (1)

where:

K

c

: proportional constant; and

Error: (actual-set point) suction pressure swing.

The presently preferred PI control algorithm implemented by the decision logic module

166

is illustrated in FIG.

11

. The routine begins at step

200

by reading the user supplied parameters K, T

i

, T

c

and S

t

. See

FIG. 6

for a description of these user supplied values. The constant K

p

is calculated as being equal to the initially supplied value K; and the constant K

i

is calculated as the product of the initially supplied constant K and the ratio T

c

/T

i

.

Next, at step

202

a decision is made whether the absolute value of the error between set point temperature and conditioned temperature (on lead

190

,

FIG. 9

) is greater than 5° F. If so, the constant K

p

is set equal to zero in step

204

. If not, the routine simply proceeds to step

206

where a new loading percent value is calculated as described by the equation in step

206

of FIG.

11

. If the load percent is greater than

100

(step

208

), then the load percent is set equal to 100% at step

210

. If the load percent is not greater than 100% but is less than 0% (step

212

) the load percent is set equal to 0% at step

214

. If the load percent is between the 0% and 100% limits, the load percent is set equal to the new load percent at step

216

.

The variable duty cycle control signal generated by the controller can take several forms.

FIGS. 12 and 13

give two examples.

FIG. 12

shows the variable duty cycle signal in which the duty cycle varies, but the frequency remains constant. In

FIG. 12

, note that the cycle time, indicated by hash marks

220

, are equally spaced. By comparison,

FIG. 13

illustrates the variable duty cycle signal wherein the frequency is also varied. In

FIG. 13

, note that the hash marks

220

are not equally spaced. Rather, the waveform exhibits regions of constant frequency, regions of increasing frequency and regions of decreasing frequency. The variable frequency illustrated in

FIG. 13

is the result of the adaptive modulation of the cycle time T

cyc

.

FIG. 14

graphically shows the benefits that the control system of the invention has in maintaining tighter temperature control and higher suction pressure with improved system efficiency. Note how the temperature curve

222

of the invention exhibits considerably less fluctuation than the corresponding temperature curve

224

of a conventional controller. Similarly, note that the pressure curve

226

of the invention has a baseline well above that of pressure curve

228

of the conventional controller. Also, the peak-to-peak fluctuation in pressure exhibited by the invention (curve

226

) is much smaller than that of the conventional controller (curve

228

).

The controller of the invention operates at a rate that is at least four times faster (typically on the order of at least eight times faster) than the thermal time constant of the load. In the presently preferred embodiment the cycle time of the variable duty cycle signal is about eight times shorter than the time constant of the load. By way of non-limiting example, the cycle time of the variable duty cycle signal might be on the order of 10 to 15 seconds, whereas the time constant of the system being cooled might be on the order of 1 to 3 minutes. The thermal time constant of a system being cooled is generally dictated by physical or thermodynamic properties of the system. Although various models can be used to describe the physical or thermodynamic response of a heating or cooling system, the following analysis will demonstrates the principle.

Modeling the Thermal Time Constant of the System Being Cooled

One can model the temperature change across the evaporator coil of a refrigeration system or heat pump as a first order system, wherein the temperature change may be modeled according to the following equation:

ΔT=ΔT

ss

[1−exp(−

t/γ

)]+

ΔT

0

exp(−τ/γ).

where:

ΔT=air temperature change across coil

ΔT

ss

=steady state air temperature change across coil

ΔT

0

=air temperature change across the coil at time zero

t=time

γ=time constant of coil.

The transient capacity of the unit can be obtained by multiplying the above equation by the air mass flow rate (m) and specific heat at constant pressure (C

p

) and integrating with respect to time.

Generally, it is the removal of the refrigerant from the evaporator that controls the time required to reach steady state operating condition, and thus the steady state temperature change across the condenser coil. If desired, the system can be modeled using two time constants, one based on the mass of the coil and another based on the time required to get the excess refrigerant from the evaporator into the rest of the system. In addition, it may also be desirable to take into account, as a further time delay, the time lag due to the large physical distance between evaporator and condenser coils in some systems.

The thermal response of the evaporator coil may be modeled by the following equation:

=½[(1

−e

1/γ1

)+(1

−e

1/γ2

)]

where:

=temperature change across coil/steady state temperature change across coil

t=time

γ

1

=time constant based on mass of coil

γ

2

=time constant based on time required to remove excess refrigerant from evaporator

In practice, the controller of the invention cycles at a rate significantly faster than conventional controllers. This is because the conventional controller cycles on and off in direct response to the comparison of actual and set-point temperatures (or pressures). In other words, the conventional controller cycles on when there is demand for cooling, and cycles off when the error between actual and set-point temperature is below a predetermined limit. Thus the on-off cycle of the conventional controller is very highly dependent on the time constant of the system being cooled.

In contrast, the controller of the invention cycles on and off at a rate dictated by calculated values that are not directly tied to the instantaneous relation between actual and set-point temperatures or pressures. Rather, the cycle time is dictated by both the cycle rate and the duty cycle of the variable duty cycle signal supplied by the controller. Notably, the point at which the controller cycles from on to off in each cycle is not necessarily the point at which the demand for cooling has been met, but rather the point dictated by the duty cycle needed to meet the demand.

Adaptive Tuning

The controller Geneva described above can be configured to perform a classic control algorithm, such as a conventional proportional-integral-derivative (PID) control algorithm. In the conventional configuration the user would typically need to set the control parameters through suitable programming. The controller may also be of an adaptive type, described here, to eliminate the need for the user to determine and program the proper control parameters.

Thus, one important advantage of the adaptive controller is its ability to perform adaptive tuning. In general, tuning involves selecting the appropriate control parameters so that the closed loop system is stable over a wide range of operating conditions, responds quickly to reduce the effect of disturbance on the control loop and does not cause excessive wear of mechanical components through continuous cycling. These are often mutually exclusive criteria, and a compromise must generally be made. In

FIG. 18

(and also

FIG. 6

) there are two basic control loops: the refrigeration control loop and the adaptive tuning loop. The refrigeration control loop is administered by control block module

102

; the adaptive loop is administered by adaptive tuning module

120

. Details of the adaptive tuning module

120

are shown in

FIGS. 15

,

16

a-

16

c

and

17

. The presently preferred adaptive tuning module uses a fuzzy logic control algorithm that will be described in connection with

FIGS. 18-20

.

Referring to

FIG. 15

, the adaptive tuning module performs basically three functions. First, it decides whether to perform adaptive tuning. This is handled by module

240

. Second, it gathers the needed parameters for performing adaptive tuning. This is handled by module

242

. Third, it calculates the adaptive gain used by the control loop. This is handled by module

244

.

Module

240

bases the decision on whether to start tuning upon two factors: the current operating state of the system and the control set point. The flowchart of

FIG. 16

a

shows the steps involved in this decision. Module

242

integrates key parameters needed for the calculations performed by module

244

. Essentially, module

242

inputs the percent loading, the temperature and pressure values and the set point temperature. It outputs the following data: S_ER (the total number of conditioned temperature and pressure data points that are within 0.5 degrees or 1 Psig of the set point value), S_Close (the total number of percent loading data points that goes to zero percent during a given sampling interval, e.g. 30 min.), S_Open (the total number of percent loading data points that goes to 100% in the sampling interval) and SSLP (a moving average or rolling average of the percent loading during the sampling interval). Module

242

is responsive to a tuning flag that is set by module

240

. Module

242

performs the integration of these key parameters when signaled to do so by the tuning flag.

FIG. 16

b

shows the steps involved in performing integration of these key parameters.

Finally, the calculation block takes the data supplied by module

242

and calculates the adaptive gain using the process illustrated in

FIG. 16

c.

The adaptive tuning module

120

will cycle through various operating states, depending on the state of a timer.

FIG. 17

is a state diagram showing how the presently preferred embodiment will function. Note that the sequence transitions from the initialization mode to either the integration mode or the no tuning mode, depending on whether the tuning flag has been set. Once in the integration mode, the system performs integration until the timer lapses (nominally 30 minutes), whereupon the calculation mode is entered. Once the calculations are completed the timer is reset and the system returns to the initialization mode.

The block diagram of the adaptive scheme is shown in FIG.

18

. There are two basic loops—The first one is the PID control loop

260

that runs every “dt” second and the second is the adaptive loop

262

that runs every “ta” second. When the control system starts, the PID control loop

260

uses a default value of gain (K) to calculate the control output. The adaptive loop

262

, checks the error e(t)

264

every “ta” seconds

266

(preferably less than 0.2 * dt seconds). At module

268

if the absolute value of error, e(t), is less than desired offset (OS), a counter Er_new is incremented. The Offset (OS) is the acceptable steady-state error (e.g. for temperature control it may be +/1° F.). This checking process continues for “tsum” seconds

270

(preferably 200 to 500 times dt seconds). After “tsum” seconds

270

, the value Er_new is converted into percentage (Er_new%

272

). The parameter Er_new%

272

indicates the percentage of sampled e(t) that was within accepted offset (OS) for “tsum” time. In other words, it is a measure of how well the control variable was controlled for past “tsum” seconds. A value of 100% means “tight” control and 0% means “poor” control. Whenever Er_new% is 100%, the gain remains substantially unchanged as it indicates tighter control. However, if Er_new happens to be between 0 and 100%, adaptive fuzzy-logic algorithm module

274

calculates a new gain (K_new

276

) that is used for next “tsum” seconds by the control algorithm module

278

.

In the preferred embodiment, there is one output and two inputs to the fuzzy-logic algorithm module

274

. The output is the new gain (K_new) calculated using the input, Er_new%, and a variable, Dir, defined as follows:

Dir=

Sign[(

ER

-new%−

ER

_old%)*(

K

_new−

K

_old)] (2)

where:

Sign stands for the sign (+ve, −ve or zero) of the term inside the bracket;

Ernew% is the percentage of e(t) that is within the offset for past “tsum” seconds;

Er_old% is the value of Er_new% in “(tsum−1)” iteration;

K_new is the gain used in “tsum” time; and

K_old is the gain in (tsum−1) time.

For example, suppose the controller starts at 0 seconds with a default value of K=10 and, ta=1 seconds, tsum=1000 seconds and OS=1. Suppose 600 e(t) data out of a possible 1000 data was within the offset. Therefore, after 1000 sec. Er_new%=60 (i.e., 600/1000*100), K_new=10. Er_old% and K_old is set to zero when the adaptive fuzzy-logic algorithm module

274

is used the first time. Plugging these numbers in Eq.(2) gives the sign of the variable “Dir” as positive. Accordingly, the inputs to the adaptive fuzzy-logic module

274

for the first iteration are respectively, Er_new%=60 and Dir=+ve.

The next step is to perform fuzzification of these inputs into fuzzy inputs by using membership functions.

Fuzzification

A membership function is a mapping between the universe of discourse (x-axis) and the grade space (y-axis). The universe of discourse is the range of possible values for the inputs or outputs. For ER_new% it is preferably from 0 to 100. The value in the grade space typically ranges from 0 to 1 and is called a fuzzy input, truth value, or a degree of membership.

FIG. 19

shows graph

300

which contains the membership functions for the input, Er_new%. Er_new% is divided into three linguistic variables—LARGE (

304

), MEDIUM (

306

) AND SMALL (

308

). For Er_new%=60, the fuzzy inputs (or degree of membership function) are −0.25 of LARGE and 0.75 of MEDIUM. The input variable “Dir” is well defined (+ve, −ve or zero) and thus does not require a membership function in this application. The next step is to create the “Truth Table” or Rule Evaluation.

Rule Evaluation

Rule evaluation takes the fuzzy inputs from the fuzzification step and the rules from the knowledge base and calculates fuzzy outputs.

FIG. 20

shows the rules as truth table. For the first column and first row, the rule is: “IF ER_new% is LARGE AND Dir is NEGATIVE THEN New Gain is NO CHANGE (NC)” (i.e. if the percentage of e(t) data that is within the offset (OS) for last “tsum” seconds is LARGE and the direction (DIR) is NEGATIVE/ZERO then do not change the existing K value (NO CHANGE)).

In the example, because ER_new% has fuzzy inputs LARGE (0.25) AND MEDIUM (0.75) with POSITIVE Dir, the rules that will be used are:

IF ER_new% is LARGE (0.25) AND Dir is POSITIVE THEN New Gain is NO CHANGE (NC=1)

IF ER_new% is MEDIUM (0.75) AND Dir is POSITIVE THEN New Gain is POSITIVE SMALL CHANGE (PSC=1.2)

Defuzzification

Finally, the defuzzification process converts the fuzzy outputs from the rule evaluation step into the final output by using Graph

310

of FIG.

21

. Graph

310

, uses the following labels =“NBC” for negative big change; “NSC” for negative small change; “NC” for no change; “PSC” for positive small change; and PBC for positive big change. The Center of Gravity or centroid method is used in the preferred embodiment for defuzzification. The output membership function for change in gain is shown in FIG.

21

.

The centroid (the Fuzzy-Logic Output) is calculated as:

Centroid = K_new \cdot [\frac{\sum μ (x) \cdot x all . x}{\sum μ (x) all . x}]

where:

(x) is the fuzzy output value for universe of discourse value x. In our example, the output (K_new) becomes

Output = 10 \cdot [\frac{0.25 (1) + 0.75 (1.2)}{0.25 + 0.75}] \approx 11.50

Once the three steps of fuzzification, rule evaluation, and defuzzification are finished and the output has been calculated, the process is repeated again for new set of Er_new%.

In the above example, after the first 1000 sec, the adaptive algorithm calculates a new gain of K_new=11.50. This new gain is used for the next 1000 sec (i.e. from t=1000 to 2000 sec in real time) by the PID control loop. At t=1001 sec, counter Er_new is set to zero to perform counting for the next 1000 seconds. At the end of another 1000 seconds (ie. at t=2000 seconds), Er_new% is calculated again.

Suppose this time, Er_new% happens to be 25. This means, by changing K from 10 to 11.5, the control became worse. Therefore, it would be better to change gain in the other direction, i.e., decrease the gain rather than increase. Thus, at t=2000 sec, Er_new%=25, Er_old%=60 (previous value of Er_new%), K_new=11.5 and K_old=10 (previous value of K). Applying Eq.(2), a negative “Dir” is obtained. With Er_new% of 25 and Dir=Negative, the fuzzy-logic calculation is performed again to calculate a new gain for the next 1000 seconds. The new value of gain is K_new=7.76 and is used from t=2000 to 3000 seconds by the PID Loop.

Suppose for the third iteration, i.e., from t=2000 to 3000 seconds, Er_new% comes out to be 95% (which represents tighter control). Performing the same fuzzy-logic operation gives the same value of K_new, and the gain remains unchanged until Er_new% again degrades.

Exemplary Applications

Both pulse width modulated (PWM) Compressors and electronic stepper regulator (ESR) Valves can be used to control evaporator temperature/pressure or evaporator cooling fluid (air or water) temperature. The former controls by modulating the refrigerant flow and the latter restricts the suction side to control the flow. Referring back to

FIG. 18

, the block diagram of the control system for such an actuator working in a refrigeration system

279

is shown. In

FIG. 18

one and preferably up to four temperatures of evaporator cooling fluid or one evaporator suction pressure (generally shown at

282

) is sampled every dt seconds. A sampling time of dt=10 seconds was found to be optimum for both the applications. After processing by the analog to digital module

284

, the sampled signal is then reduced to one number by taking the average or the minimum or the maximum of the four temperatures depending on the system configuration or the user preference at module

286

. Typically, in a single actuator (PWM/ESR) systems where the complete evaporator coil goes into defrost at one time, averaging of control signal is preferred. In a multiple evaporator-single actuator system where defrost of evaporator coils does not occur at the same time, minimum is the preferred mode. The value obtained after avg/min/max is called conditioned signal. At comparison module

288

this is compared with the desired set point to calculate the error, e(t).

The control algorithm used in the loop is a Proportion-integral (PI) control technique (PID). The PI algorithm calculates the valve position (0-100%) in case of ESR or calculates the percentage loading (0 to 100%) in case of PWM compressor. A typical integral reset time, Ti, for both the actuators is 60 seconds. The gain is tuned adaptively by the adaptive loop. The adaptive algorithm is turned off in the preferred embodiment whenever: the system is in defrost; is going through pull-down; there a big set point change; sensor failure has been detected; or any other system failure is detected.

Consequently, the adaptive algorithm is typically used when the system is working under normal mode. The time “ta” preferably used is about 1 seconds and “tsum” is about 1800 seconds (30 minutes). Diagnostics Related to PWM Compressor/ESR Valves:

Referring to

FIG. 22

, a discharge cooling fluid temperature sensor

312

(Ta), an evaporator coil inlet temperature sensor

314

(Ti) and an evaporator coil exit temperature sensor

316

(To) can provide diagnostic features for the evaporator control using PWM/ESR. The Inlet temperature sensor

314

can be anywhere in evaporator coil

318

. However, the preferred location is about one third of the total evaporator length from the evaporator coil distributor

320

.

Using these three temperature sensors, system learning can be performed that can be used for diagnostics. For example, diagnostics can be performed for ESR/PWM when it is used in a single evaporator along with an expansion valve. In this example, the following variables are tracked every “tsum” second in the adaptive loop. The variables can be integrated just after ER_new integration is done in the adaptive loop.

N-Close: Number of times Valve position /PWM loading was 0%.

N-Open: Number of times Valve position/PWM loading was 100%.

MAVP: The moving average of the Valve position /PWM loading for “tsum” seconds.

SSLP: The steady-state Valve position /PWM loading is set equal to MAVP if for the “tsum” duration ER_new% is greater than 50%.

dT: Moving average of the difference between Ta and Ti (Ta−Ti).

SH: Moving average of the difference between To and Ti (To−Ti) in the said duration. This is approximately the evaporator superheat.

N_FL: Number of times To was less than Ti during the said duration, i.e., “tsum” seconds. This number will indicate how much the expansion valve is flooding the evaporator.

In addition, Pull-down time after defrost, tpd, is also learnt. Based on these variables, the following diagnostics are performed: temperature sensor failure; degraded expansion valve; degraded ESR valve/PWM Compressor; oversized ESR/PWM; undersized ESR/PWM; and no air flow.

Temperature Sensor Failure

Failures of temperature sensors are detected by checking whether the temperature reading falls within the expected range. If PWM/ESR is controlled using Ta as the control variable, then when it fails, the control is done as follows. The above said actuator is controlled based on Ti, or the Ta values are estimated using the learned dT (i.e., add dT to Ti value to estimate Ta). During pull down, the valve/PWM can be set to full-open/load for the learned pull-down time (tpd). If Ti also fails at the same time or is not available, the actuator is opened 100% during pull down time and then set to steady-state loading percent (SSLP) after pull-down-time. An alarm is sent to the supervisor upon such a condition.

Degraded Expansion Valve

If an expansion valve sticks or is off-tuned or is undersized/oversized, the following combinations of the tracked variable can be used to diagnose such problems. N_FL>50% and ER_new%>10% indicate the expansion valve is stuck open or is off-tuned or may be even oversized and thus is flooding the evaporator coil. An alarm is sent upon such a condition. Moreover, SH>20 and N_FL=0% indicate an off-tuned expansion valve or an undersized valve or the valve is stuck closed.

Degraded ESR Valve/PWM Compressor

A degraded ESR is one that misses steps or is stuck. A degraded PWM Compressor is one whose solenoid is stuck closed or stuck open. These problems are detected in a configuration where defrost is performed by setting the ESR/PWM to 0%. The problem is detected as follows.

If ER_new%>50% before defrost and during defrost Ti<32□ F. and SH>5□ F., then the valve is determined to be missing steps. Accordingly, the valve is closed by another 100% and if Ti and SH remain the same then this is highly indicative that the valve is stuck.

If ER_new%=0 and N_Close is 100% and Ti<32 F. and SH>5 F. then PWM/ESR is determined to be stuck open. If ER_new%=0 and N_Open is 100% and Ti>32 F. and SH>5 F. then PWM/ESR is determined to be stuck closed.

Over-sized ESR/PWM

If N_Close>90% and 30%<ER_new%<100%, then an alarm is sent for oversized valve/PWM Compressor.

Under-sized ESR/PWM

If N_Open>90% and ER_new%=0 and SH>5, then an alarm is sent for undersized valve/PWM Compressor.

No Air Flow

If N_Open=100%, ER_new%=0, SH<5 F. and Ti<25 F and N_FL>50%, then either the air is blocked or the fans are not working properly.

Additionally, these diagnostic strategies can also be applied to an electronic expansion valve controller.

The embodiments which have been set forth above were for the purpose of illustration and were not intended to limit the invention. It will be appreciated by those skilled in the art that various changes and modifications may be made to the embodiments discussed in this specification without departing from the spirit and scope of the invention as defined by the appended claims.

APPENDIX

Pseudocode for Performing Signal Conditioning

Repeat the following every Ts Seconds:

Read User Inputs:

Sampling Time (Ts)

Control Type (P or T)

Sensor Mode (Avg/Min/Max)

Perform Analog to Digital Conversion (ADC)

on all (four) Temp. Sensor Channels

output data as Counts

Digitally Filter Counts

Ynew=0.75 Yold+0.25 Counts

output data as Filtered Counts

Convert Filtered Counts to Deg F.

Test if at least one Sensor is within normal operating limits

e.g. within −40 and +90 F.

If none are within limit—Set Sensor Alarm to TRUE

Else Perform Avg/Min/Max operation based on Sensor Mode

If Control Type is NOT a T/P Control Type

Then End Signal Conditioning Routine (until next Ts cycle)

Else (Control Type is T/P) Do the Following:

Perform ADC on Pressure Sensor Channel

output data as Counts

Digitally Filter Counts

Ynew=0.75 Yold+0.25 Counts

output data as Filtered Counts

Convert Filtered Counts to Psig

Test if pressure Sensor is within normal operating limits

e.g. within 0 and +200

If not within limit:

Set dP=dP Set Pt.

Else:

Calculate dP=Pmax-Pmin

Set Sensor Alarm to Conditioned T/dP

End Signal Conditioning Routine (until next Ts cycle)

Number	Name	Date	Kind
4651535	Alsenz	Mar 1987	A
5035119	Alsenz	Jul 1991	A
5241833	Ohkoshi	Sep 1993	A
5243829	Bessler	Sep 1993	A

	Number	Date	Country
Parent	08/486118	Jun 1995	US
Child	08/939779		US

Adaptive control for a refrigeration system using pulse width modulated duty cycle scroll compressor

Information

Patent Number

Date Filed

Date Issued

Inventors

Original Assignees

Examiners

Agents

CPC

US Classifications

Field of Search

US

International Classifications

Abstract

Description

Claims

CROSS-REFERENCE TO RELATED APPLICATION

US Referenced Citations (4)

Continuation in Parts (1)