Abstract:
The control system employs multiple, distributed architechure, processors in the indoor unit, the outdoor unit and optionally the thermostat unit. The overall system control functions are distributed across the processors, which communicate with one another over a communications link, employing a messaging system which allows sensor values to be obtained locally and transmitted to the other processors. The processors communicate to automatically refresh or update information during times when it will not degrade performance of critical functions.

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application is a continuation-in-part of U.S. patent application Ser. No. 07/928,892, filed Aug. 12, 1992, U.S. Pat. No. 5,311,748, entitled &#34;Control System for Heat Pump Having Decoupled Sensor Arrangement&#34;; this application is also a continuation-in-part of U.S. patent application Ser. No. 07/961,139, filed Oct. 14, 1992, U.S. Pat. No. 5,303,561 entitled &#34;Control System for Heat Pump Having Humidity Responsive Variable Speed Fan.&#34; 
    
    
     BACKGROUND AND SUMMARY OF THE INVENTION 
     The present invention relates generally to electronic control systems for air conditioners, heat pumps and refrigeration equipment. More particularly, the invention relates to a control system using a multiple microprocessor, distributed architecture configuration for controlling refrigerant flow, fan speed and other components, based on environmental measurements. 
     The heating and cooling industry has been striving for some time to improve the efficiency of air-conditioning and heat pump systems. At the same time, there has been trend of adding new features and systems to increase occupant comfort and ease of use. 
     System efficiency is a multifaceted concept, with a number of different ways of measuring and evaluating efficiency. Currently in the United States the parameters by which efficiency may be measured and evaluated are established by the Air-Conditioning and Refrigeration Institute (ARI). To allow different makes and models of air-conditioning and heat pump equipment to be compared for efficiency, ASHRAE has promulgated the efficiency measurement standard set forth in ARI 210-81. By those standards, for example, efficiency is measured at the unit&#39;s full rated output, i.e., with the unit operating at maximum heating or cooling effort. While this method of rating efficiency does allow units of different manufacturers to be compared on a somewhat common ground, in day to day operation, most heating and cooling equipment cannot be expected to operate continuously at full rated output. In fact, we have found that operating heating and cooling equipment at all times at full rated output actually degrades system efficiency. At many times during the day, and in some regions of the country at nearly all times, optimal efficiency is achieved at well below the maximum rated output. In these regions, energy could actually be saved if the heating and cooling equipment were scaled back to operate below peak output. 
     To do so, however, is not simply a matter of adjusting the thermostat up or down a few degrees, or scaling back the amount of refrigerant pumped through the system. Heating and cooling systems which have been designed to operate at optimum efficiency during full rated output will not necessarily operate at optimum efficiency when scaled back to less than full rated output. Optimal efficiency is ordinarily fine-tuned into the system during the design and manufacturing process, based on a wide variety of different physical and thermodynamic constraints which are not readily alterable after the system is built. Thus it has heretofore been impractical and uneconomical to provide heating and cooling systems with the ability to optimize efficiency over the normally encountered range of operating limits. Because ARI standards base efficiency ratings on full rated output, heating and cooling systems are understandably designed to provide optimum efficiency at full rated output. In this way the systems will compare favorably with their competition. The result, of course, is that systems, which rarely if ever need to operate at full rated output, rarely or never achieve optimal efficiency. 
     Compounding the efficiency problem is the issue of comfort. To a considerable extent, relative humidity affects comfort. When conditions are dry one feels comfortable at higher temperatures than when conditions are humid. In humid weather it is often necessary for the building occupant to lower the thermostat below the normal setting just to be comfortable. One benefit of air-conditioning systems is that they remove moisture from the air being cooled. 
     Surprisingly, however, many highly efficient air-conditioning systems remove less moisture from the air than older, less efficient ones do. This is due to the fact that a highly efficient air-conditioning system nominally operates at higher evaporator coil temperature, which removes less moisture from the air passing over it. The net effect is that sometimes a highly efficient air-condition system ends up consuming more power than an inefficient one, because the user may set the thermostat lower in order to feel comfortable. Present-day air-conditioning systems have not adequately addressed this problem. 
     Another area where improvements are wanted is in the manner in which refrigerant is metered through the system. The refrigerant cools by evaporation in a heat exchanger commonly called an evaporator coil. The refrigerant is metered to the evaporator coil through an orifice sometimes called an expansion valve. Ideally a refrigeration system should meter just enough refrigerant into the evaporation coil, so that the refrigerant extracts heat throughout the entire length of the coil as it evaporates. Due to the changing dynamics of the system, changes in thermostat setting, changes in load from sun, wind and so forth, the optimal performance is not always easy to achieve. 
     For example, when the temperature surrounding the evaporator coil is high, refrigerant is rapidly converted from the liquid phase to the gaseous phase and there may not be enough refrigerant in the liquid phase to fill the entire evaporator coil. When this occurs, efficiency suffers, since, in effect, some of the evaporator coil is being wasted. On the other hand, if the temperature surrounding the evaporator coil is low, there may not be enough heat present to cause all of the refrigerant to evaporate from the liquid phase to the gaseous phase. When this occurs, liquid refrigerant may enter the compressor, degrading efficiency and possibly damaging the compressor. Conventional refrigeration systems have employed a number of different control schemes for metering the refrigerant into the evaporator coil. Although purportedly successful to some degree, there still remains a great deal of room for improvement. 
     With all of these diverse areas for improvement, it would seem desirable to employ microprocessor technology. However, use of microprocessor technology presents its own set of problems, in that microprocessor-based systems tend to employ numerous sensors (with numerous connecting wires) and tend to be very feature-oriented. Seemingly, consumers always want all of the latest features, and thus equipment manufacturers find themselves continually upgrading their product line. When new features are added, although implemented largely in software, frequently the existing hardware will also have to be replaced, at considerable expense. 
     The present invention takes a different approach to microprocessor-based control systems, in which a highly modular, multiple microprocessor system is provided. The system employs a plurality of processors (e.g. one in the outdoor unit, one in the indoor unit and optionally one in a thermostat unit) and a communications link over which the processors communicate, using a messaging system designed to allow division of labor between processors. The communications link eliminates the need to run a bundle of wires between indoor and outdoor units even though a multiplicity of sensor can be utilized. Moreover, the distributed architecture allows operating functions to be distributed over the plurality of microprocessors so that a portion of the system (such as the indoor unit, or the outdoor unit, or the thermostat) can be upgraded without the need to upgrade the other portions. Also, if a malfunction or component failure occurs in one of the units, the remaining units can in many cases still continue to operate. 
     The present invention addresses the above efficiency problems in a very effective and inexpensive way. The system employs a variable speed indoor fan which adjusts the airflow across the indoor coil or heat exchanger based on environmental measurements such as outdoor air temperature. The airflow rate across the indoor heat exchanger affects the rate at which heat is extracted (in the cooling mode) or injected (in the heating mode), which in turn affects the superheat within the refrigeration system. The microprocessor-based control system selects the appropriate indoor airflow (fan speed) to place the system in an optimal efficiency range. 
     To precisely control refrigerant flow, the system provides a microprocessor-based control system and digitally controlled valve. Precise metering of refrigerant is made possible by a decoupled sensing arrangement which is virtually immune from previously troublesome errors caused by changing system dynamics. The system is able to maintain efficient operation at low temperature levels not heretofore readily attained. 
     The present invention addresses the humidity problem by controlling the speed of the indoor fan to cause a slower airflow when conditions are humid, in order to remove a greater amount of moisture from the air. More specifically, the microprocessor-based control system determines the indoor fan speed to stay within the boundaries of the cooling mode comfort envelope defined by the American Society of Heating and Refrigeration Engineers (ASHRAE), thereby correlating room temperature and humidity. Furthermore, the two-way communication characteristic provided by this distributed architecture approach provides the system designer access to system related information (refer to FIG. 4) in the design of diagnostic functions that enhance system reliability--a feature that may not be readily available using the conventional system design approach. 
    
    
     For a more complete understanding of the invention, its objects and advantages, reference may be had to the following specification and to the accompanying drawings. 
     BRIEF DESCRIPTION OF THE DRAWINGS 
     FIG. 1 is a schematic representation of the basic refrigeration cycle; 
     FIG. 2 is a schematic diagram of a preferred embodiment of the refrigeration control system of the invention, illustrated in a heat pump application; 
     FIG. 3 is a block diagram of the presently preferred microprocessor based system; 
     FIG. 4 is a block diagram showing data flow between room unit (thermostat) and indoor unit; 
     FIG. 5 is a block diagram showing data flow between indoor unit and outdoor unit; 
     FIG. 6 is a software block diagram of the preferred software embodiment; 
     FIGS. 7a and 7b are graphs depicting the relationship between compressor superheat and compressor discharge temperature, illustrating the minimum stable superheat and unstable regions; 
     FIG. 8 is a diagrammatic view illustrating the decoupled sensor arrangement utilized by the invention; 
     FIG. 9 is a schematic representation similar to that of FIG. 8, representing the prior art temperature sensor arrangement; 
     FIG. 10 is a cross-sectional view showing a preferred manner of positioning the ambient temperature sensor of the invention to reduce the effects of radiant heating; 
     FIG. 11 is a graph depicting the relationship between compressor discharge temperature and compressor superheat for various outdoor temperatures and the recommended operating range for superheat for maximum system efficiency; 
     FIG. 12 is a graph depicting the relationship between compressor discharge temperature and outdoor air temperature for the COOLING mode; 
     FIG. 13 is a graph depicting the relationship between compressor discharge temperature and indoor air temperature for the HEATING mode; 
     FIG. 14 is a graph illustrating the presently preferred valve setting update rate as a function of outdoor temperature; 
     FIG. 15 is a graph depicting the presently preferred manner in which valve setting is changed in steps or increments based on temperature error; 
     FIG. 16 is a graph depicting the relationship between indoor airflow as a percent of maximum and outdoor temperature in the HEATING mode; 
     FIG. 17 is a graph depicting the relationship between indoor airflow as a percent of maximum and outdoor temperature in the COOLING mode; 
     FIGS. 18a and 18b are graphs depicting the temperature-humidity comfort envelope; and 
     FIGS. 19 and 20 are flowchart diagrams illustrating the presently preferred manner of controlling indoor airflow. 
    
    
     DESCRIPTION OF THE PREFERRED EMBODIMENT 
     The present invention provides an apparatus and method for optimizing the refrigeration cycle employed in air conditioners, heat pumps and refrigeration equipment. The system employs a microprocessor-based control system with a unique complement of sensors and an electronically controlled stepper motor expansion valve to optimally control the system superheat. To illustrate the principles of the invention a heat pump system capable of providing both heating and cooling will be described. A heat pump system of the type described might be suitable for heating and cooling a commercial or residential building, although the principles of the invention are not limited to commercial and residential heating and cooling and are applicable to all pumped heat transfer systems. 
     Refrigeration Cycle Overview 
     Before giving a detailed description of the presently preferred embodiment, a brief review of the refrigeration cycle will be given. That cycle will be described in connection with a basic cooling system illustrated schematically in FIG. 1. 
     Basically the refrigeration cycle uses the cooling effect of evaporation to lower the temperature or &#34;air-condition&#34; the surroundings. This is accomplished by releasing a refrigerant under pressure (usually in the liquid phase) into a low pressure region to cause the refrigerant to expand into a low temperature mixture of gas and liquid. This refrigerant mixture now in the evaporator coil 10 is exposed to the high temperature ambient air of the region desired to be cooled. Evaporation of liquid refrigerant to gas absorbs heat from circulating air and thereby cooling it. A fan or blower is often used to force the ambient air into more rapid contact with the cool evaporator coil to increase heat transfer. The refrigerant is able to evaporate or boil in the evaporator coil because of the lower pressure. In order to make the transition from the liquid to the gaseous phase the latent heat of vaporization must be added. This heat is, of course, extracted from the ambient air environment surrounding the evaporator coil. 
     The flow of refrigerant in the high pressure (liquid) state into the low pressure evaporator is usually metered by a restricted orifice or valve commonly called an expansion valve 12. There are a wide variety of different types of expansion devices in use today, ranging from simple nonadjustable capillary tubes to electrically adjustable valves such as pulse width modulated valves. 
     Were there an endless supply of refrigerant under pressure available to be expanded in the evaporator coil and if such refrigerant could then be safely exhausted to atmosphere, a refrigeration system would require only an evaporator coil in order to function. However, since there are few convenient sources of refrigerant under pressure which can be used and then exhausted to atmosphere, the basic refrigeration system also includes a recovery system which gathers the refrigerant at the outlet of the evaporator coil, compresses it back into a high pressure state and condenses it into a liquid phase to be used once again. 
     Thus the basic refrigeration cycle routes the gaseous refrigerant from the outlet of the evaporator coil to a compressor 14 which compresses the gaseous refrigerant, placing it at a substantially higher pressure, and raising its temperature. The high pressure, high temperature refrigerant, still in the gaseous phase, is then fed to a heat removal system where the high pressure refrigerant gas is allowed to condense into the liquid phase. 
     Ordinarily this heat removal system comprises another coil called the condenser coil 16. By locating the condenser coil in an ambient condition at a temperature lower than the temperature of the pressurized refrigerant, heat is extracted from the refrigerant, cooling it sufficiently to cause a transition to the liquid phase. From the outlet of the condenser coil, the liquid refrigerant may be stored in an optional liquid receiver tank 18 and thereafter recycled back to the expansion valve to begin the cycle once again. 
     The expansion of refrigerant from the liquid phase to the gaseous phase in the evaporator coil does not occur instantaneously. If one were to examine the interior of an evaporator coil 10 when the refrigeration cycle is operating, one would find at the inlet end (adjacent the expansion valve) some refrigerant would be in the gaseous phase and some would be in the liquid phase. As the refrigerant flows through the coil, more and more of the liquid will boil or evaporate into the gaseous phase, leaving less and less refrigerant in the liquid phase. 
     At some point (possibly within the coil or possibly beyond the coil outlet in the line leading to the compressor) the refrigerant will have finally completely evaporated to the gaseous phase. This point is called the dry out point. The dry out point is important because it marks the point at which the refrigerant can be elevated to a temperature above the saturation temperature or boiling temperature. This is called superheating. 
     To understand superheating one must recognize that in an enclosed system, such as in the evaporator coil, at constant pressure a boiling liquid will remain at a constant boiling temperature so long as there is some liquid remaining in the system. The boiling temperature is, of course, pressure dependent as dictated by thermodynamic laws. However, once all of the liquid refrigerant has boiled away, the gaseous refrigerant is now potentially able to absorb additional heat to increase in temperature above the boiling temperature. 
     When the gaseous refrigerant is in contact with liquid refrigerant, it is commonly said to be in the saturated state. In the saturated state the temperature of the gaseous refrigerant is the same as boiling temperature. Once all of the liquid has boiled away and if additional heat is added to elevate the temperature of the vapor above the boiling temperature, then the vapor is said to be superheated. The term superheat is used to denote the temperature difference between the vapor temperature and the saturation temperature (boiling point temperature). 
     By measuring the superheat of the refrigerant exiting the evaporator coil it is possible to acquire some understanding of how efficiently the refrigeration system is functioning. For example, if no superheat is achieved, then one can infer that the refrigerant is not fully evaporating in the evaporator coil. This, high concentration of liquid phase refrigerant in the outlet of the evaporator implies that liquid refrigerant could be fed to the compressor, a condition which greatly deteriorates compressor efficiency and can even cause compressor damage. 
     On the other hand, an excessively high superheat value implies that the evaporator coil could have accommodated more refrigerant, again an indicator of reduced efficiency. A design objective of a refrigeration control system is to meter precisely the right amount of refrigerant so that the entire length of the evaporator coil is used without flooding the coil with excessive refrigerant that may get passed to the compressor as liquid. 
     Heat Pump System Components 
     Having reviewed the basic principles of the refrigeration cycle, a presently preferred embodiment of the invention will be given. Although the invention can take many forms, it will be described in the heat pump system of FIG. 2. With reference to FIG. 2, the heat pump system is depicted generally at 20. The system includes an indoor unit 22 and an outdoor unit 24. The indoor unit includes an indoor coil or heat exchanger 26 and an indoor fan 28. The indoor fan is preferably driven by a variable speed motor 30. The indoor fan and coil are enclosed in a suitable cabinet so that the fan forces ambient indoor air across the indoor coil at a rate determined by the speed of the variable speed motor. 
     The outdoor unit includes an outdoor coil or heat exchanger 32 and an outdoor fan 34 driven by suitable motor 36. Preferably the outdoor unit comprises a protective housing which encases the outdoor coil and outdoor fan so that the fan will draw ambient outdoor air across the outdoor coil to improve heat transfer. The outdoor unit also houses compressor 38. 
     The system illustrated in FIG. 2 is a so-called &#34;heat pump&#34; system because it can be used for both cooling and heating, by simply reversing the function of the indoor coil and the outdoor coil. This is done with a four-way reversing valve 40. Referring to FIG. 2, when the four-way valve is set to the COOLING position (shown), the indoor coil functions as the evaporator coil and the outdoor coil functions as the condenser coil. When the four-way valve is switched to the HEATING position (the alternate position), the function of the coils is reversed. The indoor coil functions as the condenser and the outdoor coil functions as the evaporator. 
     In addition to the indoor and outdoor coils and the compressor, the present system also uses an electronically controllable expansion valve 42. In the presently preferred embodiment the expansion valve is a continuously variable (or incrementally variable) stepper motor valve which can be adjusted electronically to a wide range of orifice sizes or valve openings, ranging from fully open to fully closed. Although it is possible to implement the control system of the invention with other types of electrically controlled valves, pulse width modulated valves being an example, the present embodiment prefers the stepper motor valve because it provides ripple-free operation. The stepper motor valve only needs to move or cycle when an orifice size adjustment is made. This may happen several times during a typical operating sequence (e.g., several times per hour). In contrast, the pulse width modulated valve cycles continuously during the entire operating sequence. 
     Details of Microprocessor-Based Hardware and Software 
     For convenience, the microprocessor-based control system used in implementing the refrigeration system of FIG. 2 is shown in FIG. 3. Specifically, the indoor unit 22, room unit or thermostat unit 23 and outdoor unit 24 are interconnected by communication link 48. Preferably communication link 48 is a four wire bus supplying AC power and ground to all three units and also providing a pair of serial communication data lines. The indoor unit 22, room unit 23 and outdoor unit 24 each have a microprocessor-based control unit which communicates with the other units over the communication link. In FIG. 3 the microprocessor-based control units 44 and 46 have thus been illustrated. The room unit 23 may also include a microprocessor-based control unit. The details have been omitted from FIG. 3 to simplify the illustration. 
     Also associated with the indoor unit 22 is the indoor fan or blower 28, which includes an integrated drive and variable speed motor 30. The presently preferred embodiment uses a motor which requires 240 VAC. Control signals and logic signals sent via communication link 48 are at 24 VAC and 5 VDC. A step-down transformer 180 is provided on the indoor unit for this purpose. Motor 30 receives its operating control instructions from and supplies status signals to the indoor control unit 46 at 24 VAC line levels over motor communication path 182. 
     The presently preferred embodiment uses electric resistance heaters to supply auxiliary heat. Indoor control unit 46 is responsible for determining when to turn the auxiliary heat on and off. This signal is supplied at 24 VAC logic levels. The indoor control unit 46 also supplies 24 VAC logic signals to control the expansion valve 42. 
     The outdoor unit also supplies and receives a number of different control signals at 24 VAC logic levels via the paths illustrated. For example, outdoor control unit 44 supplies the ON/OFF signal to compressor relay K1. This relay in turn energizes the compressor controller 190. The outdoor control unit 44 similarly supplies ON/OFF control signals to the heat/cool relay K2, which switches the reversing valve 40 as described above to place the system in either the HEATING or COOLING mode. The outdoor control unit supplies logic signals to the fan ON/OFF relay K3 and the fan speed relay K4. These relays in turn control the outdoor fan motor 36. In the presently preferred embodiment the outdoor fan motor 36 is a two speed motor. Outdoor control unit 44 also receives logic level data from the outdoor sensors, such as first temperature sensor 54, which measures the discharge temperature of the compressor, second temperature sensor 56, which measures the ambient or environment air temperature and coil temperature sensor 55, which measures the temperature of outdoor coil 32 (FIG. 2). 
     Referring to FIGS. 4 and 5, the microprocessor-based control units 44, 45 and 46 are shown in somewhat greater detail to show the types of signals which are communicated between the units during operation. In FIG. 4 the room unit control unit 45 is shown together with its associated keypad 192 and display 194. The keypad is coupled to the room unit control unit 45 via an eight line parallel data bus. The display is similarly coupled via a 36 line parallel data bus. Note the indoor temperature sensor 60 and indoor humidity sensor 62 are also coupled to the room unit control unit 45. Note, also that a humidity sensor 196 may also be employed in the outdoor unit (see FIG. 5) so that outdoor humidity can be measured. To protect the outdoor unit, a pressure cutout 198 and a temperature cutout 200 are provided. 
     In reviewing FIGS. 4 and 5 it should be kept in mind that the presently preferred interunit communication is done via the serial communication link 48. It is the nature of serial communication that the communication line is shared by all units and thus only one message is communicated at a time. Thus in FIGS. 4 and 5, the itemized data flow between room unit control unit 45 and indoor control unit 46 and between outdoor control unit 44 and indoor control unit 46 are shown to illustrate the type of information which is passed back and forth during system operation. Although FIGS. 4 and 5 illustrate this transfer of data as a parallel operation, and it could be implemented in a parallel communication link, the present embodiment prefers a serial ink. 
     To transfer this much information back and forth during system operation the microprocessor-based systems send and acknowledge messages in a serial fashion according to a predefined protocol. Each of the three units has sufficient random access memory, and in some cases nonvolatile memory, in which to store the various parameters which comprise the information communicated between them. Thus, for example, the outdoor humidity measured by sensor 196 is stored in the random access memory of outdoor control unit 44, to be thereafter communicated to indoor control unit 46 where it is stored in the random access memory of control unit 46. Thereafter, the same outdoor humidity data may be communicated to the room unit where it is stored in the room unit control unit&#39;s random access memory. The building occupant could then, if desired, enter the appropriate keypad function to request a display of the outdoor humidity. This is but one example of how data flows from module to module. 
     By using interim storage in each of the modules, each module is able to handle its own system critical tasks with high priority, attending to the noncritical tasks only when there is time. Thus the building occupant requesting a display of outdoor humidity would not cause the outdoor module to neglect its task of controlling the compressor and likewise would not cause the indoor unit to neglect its task of regulating the fan speed. The building occupant requesting and outdoor humidity display would see a humidity value displayed instantly, although the value displayed would be the value obtained during the most recent data refresh sequence. 
     Tables I and II below illustrate the presently preferred communication protocol by which the three control units communicate. Communication between the indoor control unit and room control unit comprise a series of status requests and status replies and parameter requests and parameter replies. Similarly, communication between the indoor control unit and the outdoor control unit comprise a series of command updates, status requests and status replies. The information given in Table I is of a general nature, showing the basic handshaking protocol employed. The protocol is implemented by a series of messages back and forth between the various units. Table II gives greater detail about the presently preferred message format. Tables I and II may be read together using the key numbers appearing in circles on both Tables. Thus a status request from the indoor control unit to the room control unit (designated by the key number 1) can be readily found in both Tables. Thus the Status Request message 1 is found in both Table I and Table II. Table II gives the presently preferred message format sequence used to identify a Status Request message. From Table II the Status Request message from indoor control unit to room control unit is 81 194 19. This is a fairly short message. Other messages, such as a Status Reply 2 are considerably longer and convey considerably more information. 
     With reference to Table II it will be seen that some messages, such as those with key numbers 1, 4, 5 and 8 are simple messages which do not convey data. The remaining messages are more complex since they convey numerical data. These more complex messages each being with two byte header (e.g., message 1 begins with the bytes 81 194) used to identify the type of message and its intended recipient. In addition, these more complex messages also end in a single byte check sum (e.g. message 3 ends in check sum 223). The bytes between the header and the check sum comprise numerical data that is being transferred by the message. In Table III a brief description of each byte of data is given. For example, in message 2 the humidity set point is indicated at 80, the humidity percent is indicated at 70, and so forth. Table III defines the abbreviations used in Table II. The numerical data values shown in Table II are merely exemplary of typical operating data. Different numerical data may obtain during system operation. 
     Some of the message bytes in Table II have been designated as FLAG data. These data typically comprise ON/OFF or TRUE/FALSE Boolean data capable of being conveyed by binary numbers. To save storage space and shorten communication time, these Boolean data are combined on a bitwise level into the various FLAG bytes. For example, in message 2, the FLAG bit in the bit 7 column designates whether manual mode or normal mode is invoked. Bit 6 conveys whether the system is in power-up mode or key press mode. Bit 3 tells whether the blower is on or off. Reference may be had to Table III for the abbreviations used in describing these FLAG bits. 
     The indoor control unit 46 functions as the master, controlling the overall system performance and much of the communication among the three control units. FIG. 6 is a software block diagram illustrating the functional software elements which comprise the present control system. Because the indoor control unit functions as the master in the presently preferred embodiment, many of the software modules of FIG. 6 reside in and are performed by the indoor control unit 46. Of course, from a system standpoint, any one of the microprocessor-based modules could serve as the master. In addition, the software functions depicted in FIG. 6 could be distributed across or shared by all three control units. Thus the specific allocation of certain functions to certain ones of the control units adopted by the present embodiment should not be viewed as a limitation upon the scope of the claims. 
     Referring to FIG. 6, at the heart of the software control system is the mainline control block 202. This block or module interfaces, either directly or indirectly, with each of the remaining blocks or modules. The mainline block 202 is supervised by the main control system block 201, which supplies the global system performance requirements. The mainline control block is responsible for effectuating the instructions of the main control system by interfacing with the other modules. For example, the power up initialization and power up diagnostics blocks are called by the mainline control block during initial power up. The mainline control block has access to a math routines block 208 which is called upon anytime math calculations take place. Similarly, mainline control block 202 has access to the clock support and triac drive block 210 and the timer update/miscellaneous control block 212, which are used anywhere timing of events is involved. 
     All main control routines are located in the master control module which is preferably located in the indoor unit. A single module contains the nonvolatile storage for all other system modules and is used for: (1) system and individual modules for configuration information, (2) values which are results of self-adjusting routines, and (3) storage of system fault conditions. 
     The system is adaptive, in that it is capable of continuously updating parameters during normal operation. At initial power up, however, the system uses preset starting parameters which are stored in a nonvolatile RAM or NOVRAM memory that is accessed via the NOVRAM control/interface block 214. The nonvolatile RAM may also store other numerical values such as constants which are not ordinarily altered during system operation. The use of nonvolatile RAM for this purpose is quite advantageous, since it allows a standardized hardware system to be custom programmed in the factory before shipment. This allows a more or less standard control module package to be used for a wide range of different heating and cooling systems by simple in factory programming. This results in a considerable cost savings. 
     Communications and message handling is provided by the communications block 216 and the control support messages block 218. These blocks implement the communication protocol described in Tables I-III. The actual details regarding the manner in which messages are placed on the communication link are handled by the bus block 220. System diagnostics are performed by block 222. If desired, the system may be programmed to perform self-tests and self-configuration. This is performed by block 224. In addition, any miscellaneous support functions which do not fit the other described categories may be called by the mainline control block. For illustration purposes, the miscellaneous support block 226 is thus shown. 
     Mainline control block 202 is further responsible for the important functions of controlling the refrigeration system components, the indoor fan and any auxiliary heating equipment. With respect to the refrigeration system components, an electronic expansion valve control block 228 is provided. This block is responsible for determining and effecting the setting of the expansion valve. Similarly, the compressor control block 230 is provided for turning the compressor on and off via relay K1. A PID algorithm implemented in block 232 provides the adaptive control by which the system updates it parameters during operation. 
     Expansion Valve Control 
     Arriving at the proper setting for the expansion valve is not as simple as it might first appear. It is customary to control the expansion valve based on the superheat condition within the evaporator coil. As explained above superheat is a measure of system efficiency and can therefore be used as a control parameter. Superheat is a differential parameter representing the temperature difference or temperature elevation of the refrigerant above the saturation or boiling temperature. Being a differential parameter, superheat requires the measurement of two values in order for it to be determined. 
     Using thermodynamic laws, superheat can be calculated if one knows both the pressure and the temperature at the outlet of the evaporator coil. Compared with temperature sensors, pressure sensors are quite expensive. Hence most conventional systems determine superheat by measuring temperature at two different points (e.g., at the inlet and at the outlet of the evaporator coil). The conventional system dispenses with the pressure sensor and makes the assumption that the pressure at both temperature measurement points is the same (often an erroneous assumption). By subtracting the two measured temperatures, a superheat value is thus derived. 
     While simple, this conventional technique is not accurate. As previously explained, the position within the evaporator coil at which dry out occurs tends to migrate unpredictably during the normal refrigeration cycle. The movement of the dry out point seriously degrades the reliability of conventional temperature-temperature measurements. 
     To illustrate, consider the case in which the dry out point occurs somewhere in the midsection of the evaporator coil, far upstream of the exit temperature sensor. In this case a high superheat value will be measured, since the exit temperature sensor will be reading the temperature of a superheated vapor which has been permitted to rise in temperature from the midsection to the exit. In contrast, if the dry out point occurs downstream of the exit temperature sensor, then no superheat will be measured because the temperature sensor would in that case be exposed to a vapor in saturation. 
     Assuming a conventional system in which the temperature sensors are at two fixed locations on the evaporator coil, e.g. at the inlet and at the exit, FIGS. 7a and 7b show how superheat varies as a function of the quantity or flow of refrigerant introduced through the expansion valve. At a low flow rate, little refrigerant is introduced, dry out occurs early and the superheat value is high. At a certain high flow dry out does not occur, the system remains saturated, and superheat is zero. Due to fluctuations in the system and movement of the dry out point, there is an unstable region, designated at A in FIG. 7a in which the superheat is unpredictable. Being unpredictable, the conventional control system is constructed to avoid the unstable region by selecting a minimum stable superheat value (T min  in FIG. 7a) which avoids the unstable region. This minimum stable superheat value might be on the order of 10° to 20°, for example. Typically the unstable region corresponds to a superheat value in the neighborhood of 5°. 
     FIG. 7b is another depiction of how the minimum stable superheat value T min  may be arrived at. FIG. 7b is a graph of compressor discharge temperature as a function of compressor superheat. Both axes are plotted in degrees Fahrenheit. The graph has been drawn using empirically determined measurements, shown as rectangular boxes or data points on FIG. 7b. At temperatures above approximately 5° superheat the plotted curve shows a substantially straight line having a gently rising slope (approximately 0.9). In contrast, at temperatures below 5° superheat the data points are quite scattered over a wide compressor discharge temperature range (compare data point m with data point n). In addition, the slope of the curve in this region is quite steep (slope approximately 15.2). Based on this empirically determined data the minimum stable superheat value T min  is established at the corner or knee where the slope changes between shallow and steep. Preferably the system is designed to operate at or slightly above T min  in order to avoid the unstable region designated at A. 
     The net effect of this unstable region is to force the conventional control system to operate at a higher target superheat value. This is done to ensure that the exit temperature sensor will not be positioned in the neighborhood of the shifting dry out point throughout the range of normal operation. Being a fixed position temperature sensor, there is no convenient way to track the shifting position of the dry out point. Thus the conventional measurement system must avoid the dry out point by establishing closed loop control at a higher superheat temperature. This is inherently inefficient, but unavoidable in the conventional approach. 
     The control system of the invention avoids this inherent unpredictability by basing its measurements on phenomena which are not readily influenced by the shifting dry out point. More specifically, the present system uses a first temperature sensor which gives an indication of the operating temperature of the system, and a second temperature sensor which gives an indication of the temperature of the environment in which the system is situated. This is illustrated in FIG. 8. For comparison purposes the conventional prior art control system is illustrated in FIG. 9. Both FIGS. 8 and 9 depict a refrigeration system 50 situated in an environment 52. Both refrigeration systems have an evaporator coil 10, expansion valve 12, compressor 14 and condenser coil 16. As diagrammatically illustrated, the first temperature sensor 54 of the invention (FIG. 8) is positioned in the refrigeration system 50, but the second temperature sensor 56 is positioned in the environment 52 so that it is effectively decoupled from the refrigeration system 50. In contrast, the temperature sensors of the conventional prior art system (FIG. 9) are both situated in the refrigeration system (i.e., on opposite ends of the evaporator coil 10). 
     As shown in FIG. 2 (and also in FIG. 8) in the presently preferred embodiment the first temperature sensor 54 is positioned to measure the discharge temperature of the pressurized refrigerant as it leaves the compressor. Other sensor placements could have been chosen, but this preferred temperature placement gives accurate repeatable results which can be in part attributed to the fact that the compressor discharge temperature fluctuations are relatively well dampened from temperature disturbances caused by any shifting of dry out point. 
     The second temperature sensor 56 is preferably positioned in an airflow path which will measure the temperature of the ambient air surrounding or forced across the condenser coil 16. The ambient air temperature provides a stable reference temperature which is, for all practical purposes, fully isolated or decoupled from any operating perturbations in the refrigeration system. 
     As a result of isolating or decoupling the temperature measurement from system, the present invention is able to operate at a lower minimum stable superheat setpoint than a conventional system. In the presently preferred embodiment, the system is configured to operate at nominal superheat of 5° F. This is well within the region designated as unstable (region A) of FIG. 7a. However, because the measured variables are decoupled from perturbations caused by shifting dryout point, the present invention is stable at this low superheat temperature. 
     Placement of this ambient air temperature sensor can be critical. In the COOLING mode, with the condenser coil located outdoors, the ambient temperature sensor should be positioned so that it will not receive direct sunlight. This may be accomplished by placing the ambient temperature sensor in a tube or enclosure which is open to airflow but shielded from the direct rays of the sun. 
     FIG. 10 illustrates one manner of shielding the ambient temperature sensor. As illustrated, the sensor 56 is positioned within tube 58 which is mounted in the housing sidewall of the outdoor unit 24. The tube shields sensor 56 from the direct rays of the sun, as shown, but is open to the atmosphere and to the interior of the housing so that ambient air will flow across sensor 56 under force of outdoor fan 34. The moving airflow produced by fan 34 virtually negates any radiant and conductive heating effects caused by heating of the housing by the sun&#39;s radiant energy. The result is an accurate measurement of the ambient air temperature which will be drawn by fan 34 across outdoor coil 32. 
     In the HEATING mode the ambient air temperature sensor measures the air temperature flowing across the indoor coil. Thus the temperature sensor 60 within the system thermostat can be used to provide this signal if desired. Usually it is not necessary to be concerned with placement of the indoor air temperature sensor, since the system thermostat is usually positioned where it will not be unduly affected by direct sunlight. Although the system thermostat will usually adequately serve as the environment sensor for HEATING mode operation, if desired a separate temperature sensor could be placed in the air return ductwork to measure the temperature of the ambient air delivered across the indoor coil 26. 
     Although the compressor discharge temperature and the ambient air temperature do not directly yield a value equal to the superheat, they do yield a value indicative of the superheat. This is so because the superheat, the difference between the refrigerant saturation temperature and the elevated vapor temperature, is ultimately dictated by the temperature within the system (e.g., the compressor discharge temperature) and the environment which is placing thermal demands on the system. Indeed, the purpose of determining the superheat temperature is to obtain an indicator for proper setting of the expansion valve orifice. It has been found that the use of one system indicating temperature and one environment indicating temperature (e.g., compressor discharge temperature and ambient air temperature) gives better results, more accurate and repeatable, than is achieved by conventional systems which rely on two system dependent temperatures. 
     In order to translate the system temperature and ambient temperature measurements into the proper expansion valve setting, a calculated algorithm or lookup table may be used. The presently preferred embodiment uses a straight line or first order equation of the form y-mx+b in order to map the measured ambient temperatures onto the compressor discharge temperature setting. This is depicted graphically in FIGS. 12 and 13. The relationship between superheat (ΔT) and the compressor discharge setting may be calculated based upon empirical values for m, the slope of the line and b, the y-axis intercept. These empirical values will vary from system to system and are determined by conducting simple temperature measurements of an exemplary system over the normal operating range of temperatures for different ambients. Ideally, data points representing the most efficient operation will be selected. FIG. 11 shows data of an exemplary system with an indication of the presently preferred or recommended operating range. Once sufficient data points are measured to construct a straight line graph, similar to that of FIGS. 12 and 13, the slope m and y-axis intercept b can be determined and thereafter used to compute the proper system temperature set point based on ambient measurements. 
     The presently preferred embodiment favors a first order calculation approach over a lookup table approach. In most implementations the stepper motor valve setting will be updated at a rate which is easily accommodated by the calculation approach. Of course, in applications where the valve must be controlled on a more frequent basis or in cases where a straight line approximation is not adequate, a lookup table approach may be preferred. 
     The valve setting, is an indirect function of the superheat (ΔT), representing a desired steady state target setting. Ordinarily the valve setting will need to be updated from time to time, but not continuously. The presently preferred control system uses a modified integral control algorithm to control the way the expansion valve setting is changed over time to respond to system demands. The valve setting is updated at a rate or frequency based on outdoor temperature. Valve settings are updated in increments or steps based on the difference or error between actual set point temperature and the calculated setting determined by the y-mx+b calculation described above. 
     FIG. 14 shows the presently preferred valve setting update rate or schedule as it varies with outdoor temperature. From the curve of FIG. 14, an inverse relationship is established between valve update rate (in minutes) and outdoor air temperature. The valve setting is updated more frequently when the outdoor temperature is high, and less frequently when the outdoor temperature is low necessitated by the response time of the system at different outdoor temperatures. 
     FIG. 15 shows the presently preferred manner in which the valve setting is changed in steps or increments based on the difference or error between actual system temperature and desired temperature. Recall that the presently preferred expansion valve is a stepper motor valve. As such, it is capable of being adjusted in discrete increments or steps. As shown by the curve in FIG. 11, when the error is small, the valve is changed in small steps (e.g. one or two steps for errors less than 10° F.). When the error is great, the valve is changed in large steps (e.g. four steps at a time for errors between 10° F. to 20° F.; eight steps at a time for errors between 20° F. to 30° F.; and twelve steps at a time for errors above 30° F.). 
     From the foregoing, it will be seen that the present invention represents a significant departure from conventional refrigeration cycle control techniques. Through the use of decoupled temperature measurements (compressor discharge temperature and ambient air temperature) the heat pump or air conditioning cycle can be controlled to a degree not previously achieved. The system can be operated in a lower minimum superheat regime, unreachable by conventional systems, and at higher efficiency than previously achieved. 
     Fan Speed Control 
     Regarding the fan speed control, the blower control block 234 receives instructions from blower feedback detection block 236, pulse width modulation output control block 238 and blower ON/OFF delay block 240. The presently preferred indoor fan motor is driven by a pulse width modulation system by which the pulse width of the motor drive signal is varied to modulate the speed. The pulse width modulation output control block 238 and blower feedback detection block 236 provide the pulse width modulated closed-loop system function. The blower ON/OFF delay block 240 alters the ON/OFF sequence of the indoor fan to optimize airflow in relation to the temperature of the heat exchanging elements. 
     In systems which use auxiliary heat, such as resistance heating elements, the strip heat control block 242 and stage control block 244 are provided. The strip heat control block is responsible essentially for activating one or more relays to turn on the auxiliary heating elements. The stage control block causes the strip heat control block to do so in stages based on the desired temperature and how much heat is required. 
     The presently preferred embodiment uses a demand defrost cycle to periodically melt accumulated frost on the coil. This is handled by the defrost control block 246. In addition, fan speed may be regulated to control humidity or to maintain system operation within the ASHRAE comfort zone. This is handled by the humidity control block 248. 
     As described above presently preferred control system is a microprocessor-based system which gathers data from various sensors and which determines the proper setting of the expansion valve based on the data gathered. More specifically, the presently preferred embodiment uses two microprocessor-based control units 46 and 44, one associated with the indoor unit and one associated with the outdoor unit, respectively. In addition, the room unit or thermostat unit 23 may also include a microprocessor-based control unit 45. Preferably, all three microprocessor-based control units are linked together via a suitable communication link 48 such as a parallel or serial communication link. The outdoor control unit 44 is largely responsible for data collection while the indoor control unit 46 is responsible for: on/off cycling of system, modulating the indoor fan speed, control of expansion valve, start/termination of demand defrost malfunction detection and performing system. 
     More specifically, microprocessor 46 modulates the indoor fan speed differently, depending on whether the system is operating in HEATING mode or in COOLING mode. In both cases the indoor fan speed is controlled as a function of outdoor temperature. In the HEATING mode the fan speed is varied over a range of 70% to 100% in four steps depicted in FIG. 16. In the cooling mode fan speed is varied from 70% to 100% in two steps, as shown in FIG. 17. Of course, it is to be understood that the presently preferred airflow-temperature relationships, depicted in FIGS. 16 and 17, may be varied without departing from the spirit of the invention. Essentially, the indoor airflow is modulated based on outdoor air temperature to achieve optimum heating and cooling efficiencies. 
     Referring to FIG. 20, the indoor airflow is modulated according to an algorithm which distinguishes between the HEATING mode and the COOLING mode. If the system is in either the COOLING mode or in the HEATING mode, a check is performed at Step 404 to determine whether the compressor is running. If not, and if the indoor fan has not been set to manual mode (Step 406), then the indoor fan will be turned off as indicated at Step 408. If the indoor fan has been set to manual mode, then the indoor fan will be commanded to operate at the built-in minimum speed as indicated at Step 416. 
     If the compressor is on, the outdoor air temperature is read at Step 410 and the indoor fan is started at Step 412. A test is performed at Step 418 to ensure that the fan is on. Thereafter, if the system is in the COOLING mode, the indoor fan speed is adjusted according to the strategy of FIG. 17. See Step 424. Alternatively, if the system is in the HEATING mode, the indoor fan speed is adjusted according to the strategy of FIG. 16. See Step 426. As previously described, the strategy of FIG. 17 is a stepwise function comprising a first step calling for a 70% airflow at temperatures below approximately 92° and calling for an airflow of 100% at temperatures above approximately 92°. The strategy of FIG. 16 is also a stepwise function calling for four steps as illustrated. In addition to the stepwise control of fan speed, the presently preferred system also employs an active humidity control routine 428. Preferably this routine is applied only during the COOLING mode. It is described below in connection with FIG. 19. 
     The reduction of fan speed from 100% to 70% saves a considerably amount of energy without substantially sacrificing cooling capacity. The electrical motor driving the indoor fan follows the cubic law. That is, when the motor speed is reduced by half, the motor power consumed is reduced by a factor of eight. Similarly, when the motor speed is reduced by 70% the power consumed is reduced to about 35% of full load power. To illustrate, if the fan motor draws 500 watts of power at full load, it will draw in the neighborhood of 171 watts when the speed is reduced to 70%. This represents a significant energy savings. Along the same lines, if the speed if further reduced to 50%, the motor would draw approximately 62 watts of power. Due to the cubic law relationship, the reduction in motor speed from 100% to 70% drops power consumption from 500 watts to about 170 watts. This represents a savings of 330 watts or about one-third of a kilowatt every hour the fan is in operation. A reduction of motor speed to 50% results in additional energy savings, but not as much as the savings afforded by the drop to 70%. That is, a reduction from 70% down to 50% will result in an additional savings of roughly 110 watts. Because of these diminishing returns, the presently preferred embodiment employs a 70% speed reduction as representing a reasonable compromise between energy savings and reduced airflow. Although the 70% reduction factor is presently preferred, it is possible to implement the invention using other reduction factors. 
     As depicted in FIG. 2, the room unit microprocessor 45 or thermostat module may include an indoor temperature sensor 60 and an indoor humidity sensor 62. The indoor temperature and humidity data are transferred to indoor microprocessor 46. The indoor temperature and humidity data is used to further control the indoor fan speed in order to stay within the boundaries of the ASHRAE comfort envelope to enhance the occupant comfort. Referring to FIGS. 18a-18b, the comfort envelope demonstrates how perceived temperature is a function of not only actual temperature but also humidity. For example, at 80% relative humidity, 75° actual room temperature feels like 78°. Because of this phenomenon, on humid days building occupants may have to lower the thermostat to feel comfortable. This, of course, raises utility bills, since more energy must be expended to achieve the lower temperature setting. 
     The present invention addresses the problem at its source. It has been found that highly efficient air-conditioning systems nominally operate at higher evaporator coil temperatures and therefore remove less moisture from the air. The present invention interactively compensates for this by modulating the indoor fan speed based on humidity. More specifically, the temperature and humidity sensors are located in the room unit or thermostat unit. They are used to modulate the indoor fan speed in such a way that the indoor temperature and relative humidity move within the Comfort Envelope defined by the American Society of Heating and Refrigeration Engineers. See FIG. 18a. The principle elements of the fan speed modulation algorithm or mechanism are shown in FIG. 19. 
     Referring to FIG. 19, active humidity control is accomplished by using the measured indoor temperature and humidity to compute the difference between the actual temperature and the apparent temperature based on the moisture content or humidity within the room. This difference is used to modulate the indoor airflow so that the system&#39;s moisture removal capability is altered. The relationship between the apparent temperature, the actual temperature and the relative humidity is shown in FIG. 18b. In the presently preferred embodiment, the indoor fan motor speed is not allowed to fall below 60% of the maximum speed whenever the indoor and outdoor temperatures are less than 75°. This is the case even if the apparent indoor temperature is outside the Comfort Envelope. This is done to protect the indoor coil from freezing. 
     Referring to FIG. 19, after determining which mode the system is in Steps 256 through 259, a five minute timed loop is entered at Step 261 and 262. During this five minute interval the fan speed alogrithm is performed. First, at Step 263, if the indoor temperture and outdoor temperature are less than 75°, the minimum fan speed is set at 70% as indicated at Step 264. Otherwise, the minimum fan speed is set at 60% as indicated at Step 265. 
     Thereafter, the actual room temperature and actual room humidity are determined from the indoor sensors at Step 266 and the apparent room temperature is determined at Step 267. This may be done by employing a look up table or other suitable calculation. The difference between actual temperature and apparent temperature is computed at Step 268 and this difference is used to set the fan speed. As illustrated at Steps 269, 270 and 271, if the difference is greater than zero the fan speed is decreased by a factor of 5%. If the difference is less than zero the fan speed is increased by a factor of 5% unless the existing fan speed is already at the maximum 100%. 
     Thus the fan speed is adjusted up or down in increments preferably in 5% increments. The speed increments or decrements are performed once during the five minute interval. 
     Modulation of the fan speed actually helps control the humidity within the room. A reduction in indoor fan speed (resulting in a lower indoor air flow) lowers the indoor coil temperature. This enhances the moisture removal function of the system during high indoor humidity conditions by causing a given volume of indoor air to spend a longer time in contact with the moisture removing indoor coil. Conversely, a higher indoor fan speed improves the heat transfer, thus raising the indoor coil temperature (resulting in a higher dew point temperature of the coil). This reduces the moisture removing capability of the coil, resulting in a higher indoor relative humidity. 
     Thus, the indoor fan speed is used to modify the indoor relative humidity both during the high and during the low room humidity conditions. This improves occupant comfort. This comfort mode is invoked by the operator selecting the &#34;Comfort Mode&#34; on the thermostat controller. If this feature is not desired the occupant can select the &#34;Economy Mode&#34; which will bypass or defeat the active humidity control logic. 
     If the humidity can be brought within the bounds of the Comfort Envelope by the above-described technique, the control system will control the temperature to the set point dictated by the thermostat setting. If, however, the humidity is outside of the Comfort Envelope the control system will attempt to achieve an apparent temperature comparable in comfort to the thermostat set point temperature at a 50% relative humidity. 
     From the foregoing, it will be seen that the present invention represents a significant departure from conventional refrigeration cycle control techniques. The microprocessor-based control system and digitally controlled humidity-responsive variable speed fan of the invention greatly improve steady state and cyclic performance and comfort, by optimizing the indoor fan speed in a controlled way. Energy is saved by allowing the system to operate at optimal efficiency even at below rated output and by allowing building occupants to feel comfortable at higher cooling temperature settings. 
     While the invention has been described with respect to a presently preferred heat pump embodiment, it will be understood that the invention can be applied to numerous forms of air conditioning systems and refrigeration systems. Accordingly, the principles of the invention are not limited to the embodiment described. The invention is capable of certain modification and change without departing from the spirit of the invention as set forth in the appended claims. 
     
                                           TABLE I__________________________________________________________________________Room Control Unit     Indoor Control Unit                  Outdoor Control Unit__________________________________________________________________________ ##STR1## ##STR2## ##STR3##     REPEAT__________________________________________________________________________ 
    
     
                                           TABLE II__________________________________________________________________________Enhanced Single Speed Message Format For Versions COPE1-27 And__________________________________________________________________________Above ##STR4## ##STR5## ##STR6## ##STR7##__________________________________________________________________________ 
    
     
                                           TABLE III__________________________________________________________________________Note: In message 2 the top set of abbreviations describe thenumbers shown when the system is in manual mode. Thelower set of abbreviations describe the numbers innormal running mode.__________________________________________________________________________HUM SET PT  Humidity Set PointHUM %       Measured Indoor Relative Humidity (times 2)INV SPD     Inverter Speed (0 = off; 200 = on)OUT FAN     Outdoor Fan Speed (0 = off; 100 = low; 200 = high)STP HT      Number of Levels of Resistance Heat OnEEV POS     Electronic Expansion Valve Open PositionBLWR SPD    Indoor Blower SpeedCHECK SUM   Sum (modulo 256) of Numbers in MessageHR          Hour of DayMIN         Minute of DayTEMP WHOLE  Measured Room Temperature Whole NumberTEMP FRACT  Measured Room Temperature Fractional PartSET PT      Thermostat Set PointNXT PRG TIME       Next Programmed Set Point Change TimeNXT PRG ST PT       Next Programmed Set PointBLWR SPD    Indoor Blower Speed in Fan ON ModeDIAG MODE   Diagnostic Mode NumberMNUL MODE   Manual ModeNORM        Normal Running ModePWR UP      Power UpKEY PRES    Key PressedCMFT        Comfort ModeECON        Economy ModePARAMETER SETUP       Parameter Setup ModeBLOWER ON   Blower in Fan ON ModeLOCKOUT OVRIDE       Compressor Lockout Time OverrideMODE        Off Cool Heat or Emer Operating ModeCONFIG      Outdoor ConfigurationMODE        Operating Mode of Outdoor UnitINV FRO     Inverter Frequency (Compressor ON or OFF)MALF        Outdoor Malfunction CodesSUCT TEMP   Suction TemperatureDSCH TEMP   Discharge TemperatureOTDR AMB    Outdoor Ambient TemperatureOTDR COIL1  Outdoor Coil Temperature #1OTDR COIL2  Outdoor Coil Temperature #2CRNT        Compressor CurrentOTDR % HUM  Outdoor Percent Relative HumidityST PT OVRIDE       Set Point OverrideDIAG        Diagnostic Mode NumberCOMP SPD    Compressor Speed (ON/OFF)OTDR AMB    Outdoor Ambient TemperatureOTDR % HUM  Outdoor Percent Relative HumidityMALFUNCTION System Operating Under a MalfunctionMODE ECHO   System Mode (Heat, etc.) Echoed to ThermostatAUX HT      Resistance Strip Heat ON/OFFDEFROST     Running Defrost RoutineLOCKOUT     Compressor in Lockout TimeEEV POS     Electronic Expansion Valve Open PositionBLWR SPD    Indoor Blower SpeedSTRP HT     Number of Strip Heat Levels ONOUT PID     PID Control Val uePWM OFF TIME       Minutes Strip Heat or Compressor OFF in CyclePWM ON TIME Minutes Strip Heat or Compressor ON in CyclePID DER TERM       Value of Derivative Term of PID ValueDIAGNOSTICS System in Diagnostics ModeDEFROST     System in Defrost ModeRU COMM ERROR       Communication Fault Between Indoor &amp;       ThermostatOD COMM ERROR       Communication Fault Between Indoor &amp; Outdoor__________________________________________________________________________