|Publication number||USRE42006 E1|
|Application number||US 10/760,173|
|Publication date||Dec 28, 2010|
|Filing date||Jan 16, 2004|
|Priority date||Jun 7, 1995|
|Also published as||CN1238674C, CN1272171A, CN1308633C, CN1607478A, CN1607478B, CN1664372A, CN1664373A, CN1664473A, CN1664474A, CN1664475A, CN1664476A, CN1920305A, CN1952813A, CN1952813B, CN100344923C, CN100432584C, CN100432585C, CN100513791C, CN100565050C, DE69833266D1, DE69833266T2, EP1025403A1, EP1025403A4, EP1025403B1, EP1489368A2, EP1489368A3, EP1489368B1, EP1598611A2, EP1598611A3, EP1598611B1, EP1598612A2, EP1598612A3, EP1598613A2, EP1598613A3, EP1598614A2, EP1598615A2, EP1598615A3, US6047557, US6393852, US6408635, US6438974, US6449972, US6467280, US6499305, US6662578, US6662583, US6679072, US7389649, US7419365, US7654098, US20010002239, US20010045097, US20010049942, US20020178737, US20030084672, US20030089119, US20030094004, US20040123612, US20060288715, US20070022771, WO1999017066A1|
|Publication number||10760173, 760173, US RE42006 E1, US RE42006E1, US-E1-RE42006, USRE42006 E1, USRE42006E1|
|Inventors||Hung M. Pham, Abtar Singh, Jean-Luc Caillat, Mark Bass|
|Original Assignee||Emerson Climate Technologies, Inc.|
|Export Citation||BiBTeX, EndNote, RefMan|
|Patent Citations (74), Non-Patent Citations (10), Referenced by (1), Classifications (71), Legal Events (2)|
|External Links: USPTO, USPTO Assignment, Espacenet|
This is a division of U.S. Ser. No. 09/886,592, filed Jun. 21, 2001, entitled “Adaptive Control For A Refrigeration System Using Pulse Width Modulated Duty Cycle Scroll Compressor;” which is a division of U.S. Ser. No. 09/524,364, filed Mar. 14, 2000 U.S. Pat. No. 6,408,635; which is a division of U.S. Ser. No. 08/939,779, filed Sep. 29, 1997, now U.S. Pat. No. 6,047,557; which is a continuation-in-part of U.S. Ser. No. 08/486,118, filed Jun. 7, 1995, now U.S. Pat. No. 5,741,120, each of which is incorporated herein by reference.
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, (hereby 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 caused 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.
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
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 1 below. A suitable load sensor such as temperature sensor 58 supplies the input signal used by the controller to determine pulse width.
The exemplary compressor 30 includes an outer shell 61 and an orbiting scroll member 64 supported on upper bearing housing 63 and drivingly connected 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
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
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
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 45c. Valve 45c 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.
A block diagram of the presently preferred compressor controller is illustrated in
Indicates Sensor Reading is not
within expected range
User configuration to indicate if
Min/Max/Avg is performed for
all temperature Sensors
Sampling Time (Ts)
Rate at which Signal condition-
ing block is executed
if controlled by only Temp. or
both Temp. & Pressure
Same as before
Generated by Adaptive Block
indicative some system problem
Steady state loading %
Whether system is in defrost
Time taken to pull down after
Gain used in PI algorithm
Integral Time (TI)
used in PI
Control Time (Tc)
used in PI
Control Set Pt. (St)
used in PI
What state the machine is
If defrost status of the case
If the defrost is from external
timer or Internal clock of
Time between defrost
Termination temperature for
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 (
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 auto-configurable. 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
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.
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
The decision logic module 166 (
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,
The variable duty cycle control signal generated by the controller can take several forms.
The controller of the invention operates at a rate that is at least four times faster (typically on (be 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 demonstrate 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:
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:
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.
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 docs not cause excessive wear of mechanical components through continuous cycling. These are often mutually exclusive criteria, and a compromise must generally be made. In
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
Finally, the calculation block takes the data supplied by module 242 and calculates the adaptive gain using the process illustrated in FIG. 16c.
The adaptive tuning module 120 will cycle through various operating states, depending on the state of a timer.
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 “light” control and 0% means “poor” control. Whenever Er_new% is 100%, the gain remains substantially unchanged as it indicates lighter 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:
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.
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.
Rule evaluation takes the fuzzy inputs from the fuzzification step and the rules from the knowledge base and calculates fuzzy outputs.
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)
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:
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.
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
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:
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.
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.
If N_Close>90% and 30%<ER_new%<100%, then an alarm is sent for oversized valve/PWM Compressor.
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.
|Cited Patent||Filing date||Publication date||Applicant||Title|
|US4152902||May 23, 1977||May 8, 1979||Lush Lawrence E||Control for refrigeration compressors|
|US4227862||Sep 19, 1978||Oct 14, 1980||Frick Company||Solid state compressor control system|
|US4463576||Sep 27, 1982||Aug 7, 1984||General Motors Corporation||Solid state clutch cycler with charge protection|
|US4494383||Feb 23, 1983||Jan 22, 1985||Mitsubishi Denki Kabushiki Kaisha||Air-conditioner for an automobile|
|US4506518||Apr 30, 1984||Mar 26, 1985||Pacific Industrial Co. Ltd.||Cooling control system and expansion valve therefor|
|US4574871||May 7, 1984||Mar 11, 1986||Parkinson David W||Heat pump monitor apparatus for fault detection in a heat pump system|
|US4575318||Aug 16, 1984||Mar 11, 1986||Sundstrand Corporation||Unloading of scroll compressors|
|US4610610||Sep 25, 1985||Sep 9, 1986||Sundstrand Corporation||Unloading of scroll compressors|
|US4612776||Apr 24, 1981||Sep 23, 1986||Alsenz Richard H||Method and apparatus for controlling capacity of a multiple-stage cooling system|
|US4634046||Apr 22, 1985||Jan 6, 1987||Yamatake-Honeywell Co. Limited||Control system using combined closed loop and duty cycle control functions|
|US4651535 *||Aug 8, 1984||Mar 24, 1987||Alsenz Richard H||Pulse controlled solenoid valve|
|US4744733||Jun 18, 1986||May 17, 1988||Sanden Corporation||Scroll type compressor with variable displacement mechanism|
|US4747756||Sep 30, 1987||May 31, 1988||Sanden Corporation||Scroll compressor with control device for variable displacement mechanism|
|US4764096||May 28, 1987||Aug 16, 1988||Matsushita Electric Industrial Co., Ltd.||Scroll compressor with clearance between scroll wraps|
|US4831832||Jun 15, 1987||May 23, 1989||Alsenz Richard H||Method and apparatus for controlling capacity of multiple compressors refrigeration system|
|US4843834||Jan 11, 1988||Jul 4, 1989||Sanden Corporation||Device for controlling capacity of variable capacity compressor|
|US4893480||Mar 11, 1988||Jan 16, 1990||Nippondenso Co., Ltd.||Refrigeration cycle control apparatus|
|US4910968||May 1, 1989||Mar 27, 1990||Hitachi, Ltd.||Refrigerating apparatus|
|US4951475||Jan 21, 1988||Aug 28, 1990||Altech Controls Corp.||Method and apparatus for controlling capacity of a multiple-stage cooling system|
|US4962648||Feb 13, 1989||Oct 16, 1990||Sanyo Electric Co., Ltd.||Refrigeration apparatus|
|US4974427||Oct 17, 1989||Dec 4, 1990||Copeland Corporation||Compressor system with demand cooling|
|US5006045||Dec 16, 1988||Apr 9, 1991||Seiko Epson Corporation||Scroll compressor with reverse rotation speed limiter|
|US5007247||Sep 25, 1989||Apr 16, 1991||Danfoss A/S||Refrigeration or heat pump installation|
|US5035119||Jun 30, 1986||Jul 30, 1991||Alsenz Richard H||Apparatus for monitoring solenoid expansion valve flow rates|
|US5059098||Jan 25, 1990||Oct 22, 1991||Kabushiki Kaisha Toyoda Jidoshokki Seisakusho||Apparatus for varying capacity of scroll type compressor|
|US5067326||Aug 23, 1990||Nov 26, 1991||Alsenz Richard H||Method and apparatus for controlling capacity of a multiple-stage cooling system|
|US5079929||Jul 18, 1990||Jan 14, 1992||Alsenz Richard H||Multi-stage refrigeration apparatus and method|
|US5088297||Sep 25, 1990||Feb 18, 1992||Hitachi, Ltd.||Air conditioning apparatus|
|US5115644||Apr 6, 1990||May 26, 1992||Alsenz Richard H||Method and apparatus for condensing and subcooling refrigerant|
|US5191643||Jun 12, 1990||Mar 2, 1993||Alsenz Richard H||Method and apparatus for refrigeration control and display|
|US5203179||Mar 4, 1992||Apr 20, 1993||Ecoair Corporation||Control system for an air conditioning/refrigeration system|
|US5241833 *||Jun 24, 1992||Sep 7, 1993||Kabushiki Kaisha Toshiba||Air conditioning apparatus|
|US5243827||Apr 21, 1992||Sep 14, 1993||Hitachi, Ltd.||Overheat preventing method for prescribed displacement type compressor and apparatus for the same|
|US5243829 *||Oct 21, 1992||Sep 14, 1993||General Electric Company||Low refrigerant charge detection using thermal expansion valve stroke measurement|
|US5259210||Dec 2, 1991||Nov 9, 1993||Sanyo Electric Co., Ltd.||Refrigerating apparatus and method of controlling refrigerating apparatus in accordance with fuzzy reasoning|
|US5265434||Aug 23, 1990||Nov 30, 1993||Alsenz Richard H||Method and apparatus for controlling capacity of a multiple-stage cooling system|
|US5282729||Jun 2, 1993||Feb 1, 1994||General Motors Corporation||Radical actuator for a de-orbiting scroll in a scroll type fluid handling machine|
|US5319943||Jan 25, 1993||Jun 14, 1994||Copeland Corporation||Frost/defrost control system for heat pump|
|US5342186||Jun 2, 1993||Aug 30, 1994||General Motors Corporation||Axial actuator for unloading an orbital scroll type fluid material handling machine|
|US5381669||Jul 21, 1993||Jan 17, 1995||Copeland Corporation||Overcharge-undercharge diagnostic system for air conditioner controller|
|US5392612||Jun 4, 1993||Feb 28, 1995||Richard H. Alsenz||Refrigeration system having a self adjusting control range|
|US5440891||Jan 26, 1994||Aug 15, 1995||Hindmon, Jr.; James O.||Fuzzy logic based controller for cooling and refrigerating systems|
|US5440894||May 5, 1993||Aug 15, 1995||Hussmann Corporation||Strategic modular commercial refrigeration|
|US5447420||Apr 13, 1994||Sep 5, 1995||Copeland Corporation||Scroll compressor with liquid injection|
|US5493867||Jun 13, 1994||Feb 27, 1996||Whirlpool Corporation||Fuzzy logic adaptive defrost control|
|US5502970||May 5, 1995||Apr 2, 1996||Copeland Corporation||Refrigeration control using fluctuating superheat|
|US5515267||Mar 1, 1993||May 7, 1996||Alsenz; Richard H.||Apparatus and method for refrigeration system control and display|
|US5555195 *||Jul 22, 1994||Sep 10, 1996||Johnson Service Company||Controller for use in an environment control network capable of storing diagnostic information|
|US5682329 *||Dec 20, 1994||Oct 28, 1997||Johnson Service Company||On-line monitoring of controllers in an environment control network|
|US5735134||May 30, 1996||Apr 7, 1998||Massachusetts Institute Of Technology||Set point optimization in vapor compression cycles|
|US5741120||Jun 7, 1995||Apr 21, 1998||Copeland Corporation||Capacity modulated scroll machine|
|US6047557 *||Sep 29, 1997||Apr 11, 2000||Copeland Corporation||Adaptive control for a refrigeration system using pulse width modulated duty cycle scroll compressor|
|US6086335||Nov 12, 1997||Jul 11, 2000||Copeland Corporation||Capacity modulated scroll machine having one or more pin members movably disposed for restricting the radius of the orbiting scroll member|
|US6438974||Feb 19, 2002||Aug 27, 2002||Copeland Corporation||Adaptive control for a refrigeration system using pulse width modulated duty cycle scroll compressor|
|US6449972||Jan 16, 2001||Sep 17, 2002||Copeland Corporation||Adaptive control for a refrigeration system using pulse width modulated duty cycle scroll compressor|
|US6467280||Jun 21, 2001||Oct 22, 2002||Copeland Corporation||Adaptive control for a refrigeration system using pulse width modulated duty cycle scroll compressor|
|US6499305||May 16, 2002||Dec 31, 2002||Copeland Corporation||Adaptive control for a refrigeration system using pulse width modulated duty cycle scroll compressor|
|US6662583||Nov 27, 2002||Dec 16, 2003||Copeland Corporation||Adaptive control for a cooling system|
|US6679072||Nov 27, 2002||Jan 20, 2004||Copeland Corporation||Diagnostic system and method for a cooling system|
|US7389649||Dec 8, 2003||Jun 24, 2008||Emerson Climate Technologies, Inc.||Cooling system with variable duty cycle capacity control|
|US7419365||Aug 31, 2006||Sep 2, 2008||Emerson Climate Technologies, Inc.||Compressor with capacity control|
|US7654098||Sep 28, 2006||Feb 2, 2010||Emerson Climate Technologies, Inc.||Cooling system with variable capacity control|
|USRE40400||Sep 30, 2003||Jun 24, 2008||Emerson Climate Technologies, Inc.||Capacity modulated scroll machine|
|USRE40554||Jun 27, 2002||Oct 28, 2008||Emerson Climate Technologies, Inc.||Capacity modulated scroll machine having one or more pin members movably disposed for restricting the radius of the orbiting scroll member|
|CN1042406A||Apr 10, 1989||May 23, 1990||日新兴业株式会社||Method of and apparatus for controlling condensing agent supply to evaporator with u-shaped tubes|
|CN1137614A||Oct 27, 1995||Dec 11, 1996||科普兰公司||Capacity modulated scroll machine|
|CN1159555A||Nov 14, 1996||Sep 17, 1997||Lg电子株式会社||Refrigerant circulation device for two evaporators adopting different evaporative temp.|
|EP0085246A1||Dec 20, 1982||Aug 10, 1983||Sanden Corporation||A control circuit for a variable displacement air conditioning compressor|
|EP0281317A1||Feb 25, 1988||Sep 7, 1988||Prestcold Limited||Refrigeration systems|
|EP0453302A1||Apr 19, 1991||Oct 23, 1991||Whitbread Plc||Refrigeration circuit including diagnostic equipment|
|EP0747597A2||Nov 1, 1995||Dec 11, 1996||Copeland Corporation||Capacity modulated scroll machine|
|EP0747598A2||Jun 6, 1996||Dec 11, 1996||Copeland Corporation||Capacity modulated scroll machine|
|GB733511A||Title not available|
|JPH07190507A||Title not available|
|1||Communication pursuant to Article 94(3) EPC received from the European Patent Office regarding Application No. 04 022920.5-2301 dated Jun. 15, 2009.|
|2||Communication pursuant to European Search Opinion regarding Application No. 05016504.2 dated Dec. 16, 2009 received from the European Patent Office.|
|3||Communication pursuant to European Search Opinion regarding Application No. 05016505.9 dated Dec. 16, 2009 received from the European Patent Office.|
|4||Extended European Search Report regarding Application No. EP 05016504 dated May 25, 2009.|
|5||First Office Action dated Jul. 4, 2008 regarding Application No. 200610128576.1, received from the Patent Office of the People's Republic of China translated by CCPIT Patent and Trademark Law Office.|
|6||First Office Action from the Mexican Institute of Industry Property dated Apr. 22, 2010 regarding Mexican Patent Application No. MX/a/2008/014196. Summary of the Office Action provided by Goodrich Riquelme Asociados law firm.|
|7||Rejection Decision regarding CN 200510064854.7 dated Feb. 6, 2009.|
|8||Second Office Action dated Apr. 17, 2009 regarding Application No. 200610128576.1 received from the Patent Office of the People's Republic of China translated by CCPIT Patent and Trademark Law Office.|
|9||Third Office Action dated Aug. 21, 2009 regarding Application No. 200610128576.1 received from the Patent Office of the People's Republic of China translated by CCPIT Patent and Trademark Law Office.|
|10||Third Office Action dated Oct. 16, 2009 regarding Application No. 200410085953.9 received from the Patent Office of the People's Republic of China (translated by Unitalen Attorneys at Law).|
|Citing Patent||Filing date||Publication date||Applicant||Title|
|US20120085512 *||Oct 7, 2011||Apr 12, 2012||Audi Ag||Vehicle cooling system|
|U.S. Classification||62/126, 62/217, 62/228.1|
|International Classification||F25B49/02, F04C27/00, A47F3/04, F25B1/04, G05D1/08, F04C18/02, F25B41/06, F25B41/04, G05D23/19, F01C1/02, F04C28/08, F25B5/02, F04C28/22, F04C28/18, F04C28/02, F04C28/28, F04C28/00, F04C28/26, F25B49/00, F04C28/06, F04B49/06|
|Cooperative Classification||F25B2700/2106, F25B2341/0653, F25B2700/21174, A47F3/04, F25B2700/21175, F04C2270/015, F25B2600/0261, F04C28/22, F04C23/008, F25B5/02, F04C18/0215, F04C28/28, F04C28/06, F04C28/00, F04C27/005, Y02B30/72, F25B41/043, F25B49/022, F25B1/04, F25B49/005, G05D23/1909, F25B2400/22, F25B2700/1933, F25B41/062, F04C28/265, F25B2700/193, F04C2270/86, F04C28/02, F04C28/08, F25B2700/2117|
|European Classification||F04C28/08, F25B49/00F, F04C28/26B, A47F3/04, F04C28/22, F25B5/02, F04C28/06, G05D23/19C2, F25B49/02B, F04C18/02B2, F04C28/02, G05D1/08B4, F25B41/04B, F04C28/28, F04C27/00C, F25B1/04, F04C28/00|
|Apr 26, 2007||AS||Assignment|
Owner name: EMERSON CLIMATE TECHNOLOGIES, INC., OHIO
Free format text: CERTIFICATE OF CONVERSION, ARTICLES OF FORMATION AND ASSIGNMENT;ASSIGNOR:COPELAND CORPORATION;REEL/FRAME:019215/0273
Effective date: 20060927
|Jun 30, 2014||FPAY||Fee payment|
Year of fee payment: 12