PATENT DOCUMENT

Abstract:
A monitoring system for monitoring operation of a refrigerant-cycle system is disclosed. A differential pressure sensor measures a pressure difference between: (i) air at a first location upstream of an evaporator of the refrigerant-cycle system; and (ii) air at a second location downstream of the evaporator of the refrigerant-cycle system. An electrical sensor measures an electrical quantity indicative of power consumption of the refrigerant-cycle system. A processing system determines whether airflow through the evaporator is restricted based on the pressure difference. The processing system calculates an efficiency of the refrigerant-cycle system based on the power consumption of the refrigerant-cycle system.

Full Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application is a continuation of U.S. Ser. No. 13/269,188, filed Oct. 7, 2011, which is a continuation of U.S. Ser. No. 11/779,203 filed Jul. 17, 2007, now U.S. Pat. No. 8,034,170, which is a continuation of U.S. Ser. No. 11/130,569, filed May 17, 2005, now U.S. Pat. No. 7,244,294, which is a continuation of U.S. Ser. No. 10/916,222, filed Aug. 11, 2004, now U.S. Pat. No. 7,275,377, the entire contents of which are hereby incorporated by reference. 
    
    
     BACKGROUND 
     1. Field of the Invention 
     The invention relates to a monitoring system for measuring the operating and efficiency of a refrigerant-cycle system, such as, for example, an air conditioning system or refrigeration system. 
     2. Description of the Related Art 
     One of the major recurring expenses in operating a home or commercial building is the cost of providing electricity to the Heating Ventilation Air Conditioning (HVAC) system. If the HVAC system is not operating at peak efficiency, then the cost of operating the system increases unnecessarily. Each pound of refrigerant circulating in the system must do its share of the work. It must absorb an amount of heat in the evaporator or cooling coil, and it must dissipate this heat—plus some that is added in the compressor—through the condenser, whether air cooled, water cooled, or evaporative cooled. The work done by each pound of the refrigerant as it goes through the evaporator is reflected by the amount of heat it picks up from the refrigeration load, chiefly when the refrigerant undergoes a change of state from a liquid to a vapor. 
     For a liquid to be able to change to a vapor, heat must be added to or absorbed in it. This is what happens in the cooling coil. The refrigerant enters the metering device as a liquid and passes through the device into the evaporator, where it absorbs heat as it evaporates into a vapor. As a vapor, it makes its way through the suction tube or pipe to the compressor. Here it is compressed from a low temperature, low pressure vapor to a high temperature, high pressure vapor; then it passes through the high pressure or discharge pipe to the condenser, where it undergoes another change of state—from a vapor to a liquid—in which state it flows out into the liquid pipe and again makes its way to the metering device for another trip through the evaporator. 
     When the refrigerant, as a liquid, leaves the condenser it may go to a receiver until it is needed in the evaporator; or it may go directly into the liquid line to the metering device and then into the evaporator coil. The liquid entering the metering device just ahead of the evaporator coil will have a certain heat content (enthalpy), which is dependent on its temperature when it enters the coil, as shown in Table 2 below and  FIGS. 2-8 . The vapor leaving the evaporator will also have a given heat content (enthalpy) according to its temperature, as shown in the refrigerant tables. 
     The difference between these two amounts of heat content is the amount of work being done by each pound of refrigerant as it passes through the evaporator and picks up heat. The amount of heat absorbed by each pound of refrigerant is known as the refrigerating effect of the system, or of the refrigerant within the system. 
     Situations that can reduce the overall efficiency of the system include, refrigerant overcharge, refrigerant undercharge, restrictions in refrigerant lines, faulty compressor, excessive load, insufficient load, undersized or dirty duct work, clogged air filters, etc. 
     Unfortunately, modern HVAC systems do not include monitoring systems to monitor the operating of the system. A modern HVAC system is typically installed, charged with refrigerant by a service technician, and then operated for months or years without further maintenance. As long as the system is putting out cold air, the building owner or home owner assumes the system is working properly. This assumption can be expensive, as the owner has no knowledge of how well the system is functioning. If the efficiency of the system deteriorates, the system may still be able to produce the desired amount of cold air, but it will have to work harder, and consume more energy, to do so. In many cases, the system owner does not have the HVAC system inspected or serviced until the efficiency has dropped so low that it can no longer cool the building. This is in part because servicing of an HVAC system requires specialized tools and knowledge that the typical building owner or home owner does not possess. Thus, the building owner or home owner must pay for an expensive service call in order to have the system evaluated. Even if the owner does pay for a service call, many HVAC service technicians do not measure system efficiency. Typically, the HVAC service technicians are trained only to make rudimentary checks of the system (e.g., refrigerant charge, output temperature), but such rudimentary checks may not uncover other factors that can cause poor system efficiency. Thus, the typical building owner, or home owner, operates the HVAC system year after year not knowing that the system may be wasting money by operating at less than peak efficiency. Moreover, inefficient use of electrical power can lead to brownouts and blackouts during heat waves or other periods of high air conditioning usage due to overloading of the electric power system (commonly referred to as the electric power grid). 
     SUMMARY OF THE INVENTION 
     These and other problems are solved by a real-time monitoring system that monitors various aspects of the operation of a refrigerant system, such as, for example, an HVAC system, a refrigerator, a cooler, a freezer, a water chiller, etc. In one embodiment, the monitoring system is configured as a retrofit system that can be installed in an existing refrigerant system. 
     In one embodiment, the system includes a processor that measures power provided to the HVAC system and that gathers data from one or more sensors and uses the sensor data to calculate a figure of merit related to the efficiency of the system. In one embodiment, the sensors include one or more of the following sensors: a suction line temperature sensor, a suction line pressure sensor, a suction line flow sensor, a hot gas line temperature sensor, a hot gas line pressure sensor, a hot gas line flow sensor, a liquid line temperature sensor, a liquid line pressure sensor, and a liquid line flow sensor. In one embodiment, the sensors include one or more of an evaporator air temperature input sensor, an evaporator air temperature output sensor, an evaporator air flow sensor, an evaporator air humidity sensor, and a differential pressure sensor. In one embodiment, the sensors include one or more of a condenser air temperature input sensor, a condenser air temperature output sensor, a condenser air flow sensor, and an evaporator air humidity sensor. In one embodiment, the sensors include one or more of an ambient air sensor and an ambient humidity sensor. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a diagram of a typical refrigerant-cycle system used in HVAC systems, refrigerators, freezers, and the like. 
         FIG. 2  is a detailed pressure-heat diagram of a typical refrigerant (R-22). 
         FIG. 3  is a pressure-heat diagram showing pressure-enthalpy changes through a refrigeration cycle. 
         FIG. 4  is a pressure-heat diagram showing pressure, heat, and temperature values for a refrigeration cycle operating with a 40° F. evaporator. 
         FIG. 5  is a pressure-heat diagram showing pressure, heat, and temperature values for a refrigeration cycle operating with a 20° F. evaporator. 
         FIG. 6  is a pressure-heat diagram showing the cycle of  FIG. 4  with a 40° F. evaporating temperature, where the condensing temperature has been increased to 120° F. 
         FIG. 7  is a pressure-heat diagram showing how subcooling by the condenser improves the refrigeration effect and the COP. 
         FIG. 8  is a pressure-heat diagram showing the cooling process in the evaporator. 
         FIG. 9A  is a block diagram of a monitoring system for monitoring the operation of the refrigerant-cycle system. 
         FIG. 9B  is a block diagram of a monitoring system for monitoring the operation of the refrigerant-cycle system, where operating data for the system is provided to a monitoring service, such as, for example, a power company or monitoring center, by using data transmission over power lines. 
         FIG. 9C  is a block diagram of a monitoring system for monitoring the operation of the refrigerant-cycle system, where operating data for the system is provided to a monitoring service, such as, for example, a power company or monitoring center, by using data transmission over a computer network. 
         FIG. 9D  is a block diagram of a monitoring system for monitoring the operation of the refrigerant-cycle system, where data regarding operation of the system is provided to a thermostat and/or to a computer system such as, for example, a site monitoring computer, a maintenance computer, a personal digital assistant, a personal computer, etc. 
         FIG. 9E  is a block diagram of a monitoring system for monitoring the operation of the refrigerant-cycle system wherein an electronically-controlled metering device is provided to allow control of the system in an energy-efficient matter. 
         FIG. 9F  is a block diagram of a thermostat control and monitoring system having a data interface device provided to the thermostat. 
         FIG. 9G  is a block diagram of a thermostat control and monitoring system having a data interface device provided to the evaporator unit. 
         FIG. 9H  is a block diagram of a thermostat control and monitoring system having a data interface device provided to the condenser unit. 
         FIG. 10  (consisting of  FIGS. 10A and 10B ) shows various sensors that can be used in connection with the system of  FIGS. 9A-H  for monitoring the operation of the refrigerant-cycle system. 
         FIG. 11  shows the temperature drop in the air through the evaporator as a function of humidity. 
         FIG. 12  shows heat capacity of a typical refrigerant-cycle system as a function of refrigerant charge. 
         FIG. 13  shows power consumed in a typical refrigerant-cycle system as a function of refrigerant charge. 
         FIG. 14  shows efficiency of a typical refrigerant-cycle system as a function of refrigerant charge. 
         FIG. 15  shows a differential-pressure sensor used to monitor an air filter in an air-handler system. 
         FIG. 16  shows a differential-pressure sensor used to monitor an air filter in an air-handler system using a wireless system to provide filter differential pressure data back to other aspects of the monitoring system. 
         FIG. 17  shows the system of  FIG. 16  implemented using a filter frame to facilitate retrofitting of existing air handler systems. 
     
    
    
     DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS 
       FIG. 1  is a diagram of a typical refrigerant-cycle system  100  used in HVAC systems, refrigerators, freezers, and the like. In the system  100 , a compressor  105  provides hot compressed refrigerant gas to a hot gas line  106 . The hot gas line  106  provides the hot gas to a condenser  107 . The condenser  107  cools the gas and condenses the gas into a liquid that is provided to a liquid line  108 . The liquid refrigerant in the liquid line  108  is provided through a metering device  109  to an evaporator  110 . The refrigerant expands back into a gas in the evaporator  110  and is provided back to the compressor  105  through a suction line  111 . A suction service valve  120  provides access to the suction line  111 . A liquid line service valve  121  provides access to the liquid line  108 . A fan  123  provides input air  124  to the evaporator  110 . The evaporator  110  cools the air and provides cooled evaporator output air  125 . An optional drier/accumulator  130  can be provided in the liquid line  108 . A fan  122  provides cooling air to the condenser  107 . 
     The metering device  109  can be any refrigerant metering device as used in the art, such as, for example, a capillary tube, a fixed orifice, a Thermostatic eXpansion Valve (TXV), an electronically controlled valve, a pulsating solenoid valve, a stepper-motor valve, a low side float, a high-side float, an automatic expansion valve, etc. A fixed metering device such as a capillary tube or fixed orifice will allow some adjustment in system capacity as the load changes. As the outdoor condensing temperature increases, more refrigerant is fed through the metering device into the evaporator  110 , increasing its capacity slightly. Conversely, as the heat load goes down, the outdoor condensing temperature goes down and less refrigerant is fed into the evaporator  110 . For a location where the load does not vary widely, fixed metering devices may float with the load well enough. However, for climates where there is a relatively greater range in temperature variation, an adjustable metering device is typically used. 
     The system  100  cools the air through the evaporator  110  by using the refrigerating effect of an expanding gas. This refrigerating effect is rated in Btu per pound of refrigerant (Btu/lb). If the total heat load is known (given in Btu/hr), one can find the total number of pounds of refrigerant that must be circulated each hour of operation of the system. This figure can be broken down further to the amount that must be circulated each minute, by dividing the amount circulated per hour by 60. 
     Because of a small orifice in the metering device  109 , when the compressed refrigerant passes from the smaller opening in the metering device  109  to the larger tubing in the evaporator  110 , a change in pressure occurs together with a change in temperature. This change in temperature occurs because of the vaporization of a small portion of the refrigerant (about 20%) and, in the process of this vaporization, the heat that is involved is taken from the remainder of the refrigerant. 
     For example, from the table of saturated R-22 in  FIG. 2 , it can be seen that the heat content of 100° F. liquid is 39.27 Btu/lb and that of 40° F. liquid is 21.42 Btu/lb; this indicates that 17.85 Btu/lb has to be removed from each pound of refrigerant entering the evaporator  110 . The latent heat of vaporization of 40° F. (17.85 Btu/lb) is 68.87 Btu/lb. This is another method of calculating the refrigerating effect, or work being done, by each pound of refrigerant under the conditions given. 
     The capacity of the compressor  105  should be such that it will remove from the evaporator  110  that amount of refrigerant which has vaporized in the evaporator  110  and in the metering device  109  in order to get the necessary work done. The compressor  105  must be able to remove and send on to the condenser  107  the same weight of refrigerant vapor, so that it can be condensed back into a liquid and so continue in the system  100  to perform additional work. 
     If the compressor  105  is unable to move this weight, some of the vapor will remain in the evaporator  110 . This, in turn, will cause an increase in pressure inside the evaporator  110 , accompanied by an increase in temperature and a decrease in the work being done by the refrigerant, and design conditions within the refrigerated space cannot be maintained. 
     A compressor  105  that is too large will withdraw the refrigerant from the evaporator  110  too rapidly, causing a lowering of the temperature inside the evaporator  110 , so that design conditions will not be maintained. 
     In order for design conditions to be maintained within a refrigeration circuit, a balance between the requirements of the evaporator  110  and the capacity of the compressor  105  is maintained. This capacity is dependent on its displacement and on its volumetric efficiency. Volumetric efficiency depends on the absolute suction and discharge pressures under which the compressor  105  is operating. 
     In one embodiment, the system  100  controls the speed of the compressor  105  to increase efficiency. In one embodiment, the system  100  controls the metering device  109  to increase efficiency. In one embodiment, the system  100  controls the speed of the fan  123  to increase efficiency. In one embodiment, the system  100  controls the speed of the fan  122  to increase efficiency. 
     In the system  100 , the refrigerant passes from the liquid stage into the vapor stage as it absorbs heat in the evaporator  110  coil. In the compressor  105  stage, the refrigerant vapor is increased in temperature and pressure, then the refrigerant gives off its heat in the condenser  107  to the ambient cooling medium, and the refrigerant vapor condenses back to its liquid state where it is ready for use again in the cycle. 
       FIG. 2  shows the pressure, heat, and temperature characteristics of this refrigerant. Enthalpy is another word for heat content. Diagrams such as  FIG. 2  are referred to as pressure-enthalpy diagrams. Detailed pressure-enthalpy diagrams can be used for the plotting of the cycle shown in  FIG. 2 , but a basic or skeleton chart as shown in  FIG. 3  is useful as a preliminary illustration of the various phases of the refrigerant circuit. There are three basic areas on the chart denoting changes in state between the saturated liquid line  301  and the saturated vapor line  302  in the center of the chart. The area to the left of the saturated liquid line  301  is the subcooled area, where the refrigerant liquid has been cooled below the boiling temperature corresponding to its pressure; whereas the area to the right of the saturated vapor line  302  is the area of superheat, where the refrigerant vapor has been heated beyond the vaporization temperature corresponding to its pressure. 
     The construction of the diagram  300  illustrates what happens to the refrigerant at the various stages within the refrigeration cycle. If the liquid vapor state and any two properties of a refrigerant are known and this point can be located on the chart, the other properties can be determined from the chart. 
     If the point is situated anywhere between the saturated liquid line  301  and vapor line  302 , the refrigerant will be in the form of a mixture of liquid and vapor. If the location is closer to the saturated liquid line  301 , the mixture will be more liquid than vapor, and a point located in the center of the area at a particular pressure would indicate a 50% liquid 50% vapor situation. 
     The change in state from a vapor to a liquid, the condensing process, occurs as the path of the cycle develops from right to left; whereas the change in state from a liquid to a vapor, the evaporating process, travels from left to right. Absolute pressure is indicated on the vertical axis at the left, and the horizontal axis indicates heat content, or enthalpy, in Btu/lb. 
     The distance between the two saturated lines  301  and  302  at a given pressure, as indicated on the heat content line, amounts to the latent heat of vaporization of the refrigerant at the given absolute pressure. The distance between the two lines of saturation  301  and  302  is not the same at all pressures, for they do not follow parallel curves. Therefore, there are variations in the latent heat of vaporization of the refrigerant, depending on the absolute pressure. There are also variations in pressure-enthalpy charts of different refrigerants and the variations depend on the various properties of the individual refrigerants. 
     There is relatively little temperature change of the condensed refrigeration liquid after it leaves the condenser  107  and travels through the liquid line  108  on its way to the expansion or metering device  109 , or in the temperature of the refrigerant vapor after it leaves the evaporator  110  and passes through the suction line  111  to the compressor  105 . 
       FIG. 4  shows the phases of the simple saturated cycle with appropriate labeling of pressures, temperatures, and heat content or enthalpy. Starting at point A on the saturated liquid where all of the refrigerant vapor at 100° F. has condensed into liquid at 100° F. and is at the inlet to the metering device, between points A and B is the expansion process as the refrigerant passes through the metering device  109 ; and the refrigerant temperature is lowered from the condensation temperature of 100° F. to the evaporating temperature of 40° F. 
     When the vertical line A-B (the expansion process) is extended downward to the bottom axis, a reading of 39.27 Btu/lb is indicated, which is the heat content of 100° F. liquid. To the left of point B at the saturated liquid line  301  is point Z, which is also at the 40° F. temperature line. Taking a vertical path downward from point Z to the heat content line, a reading of 21.42 Btu/lb is indicated, which is the heat content of 40° F. liquid. 
     The horizontal line between points B and C indicates the vaporization process in the evaporator  110 , where the 40° F. liquid absorbs enough heat to completely vaporize the refrigerant. Point C is at the saturated vapor line, indicating that the refrigerant has completely vaporized and is ready for the compression process. A line drawn vertically downward to where it joins the enthalpy line indicates that the heat content, shown at h c  is 108.14 Btu/lb, and the difference between h a  and h c  is 68.87 Btu/lb, which is the refrigerating effect, as shown in an earlier example. 
     The difference between points h z  and h c  on the enthalpy line amounts to 86.72 Btu/lb, which is the latent heat of vaporization of 1 lb of R-22 at 40° F. This amount would also exhibit the refrigerating effect, but some of the refrigerant at 100° F. must evaporate or vaporize in order that the remaining portion of each pound of R-22 can be lowered in temperature from 100° F. to 40° F. 
     All refrigerants exhibit properties of volume, temperature, pressure, enthalpy or heat content, and entropy when in a gaseous state. Entropy is defined as the degree of disorder of the molecules that make up the refrigerant. In refrigeration, entropy is the ratio of the heat content of the gas to its absolute temperature in degrees Rankin. 
     The pressure-enthalpy chart plots the line of constant entropy, which stays the same provided that the gas is compressed and no outside heat is added or taken away. When the entropy is constant, the compression process is called adiabatic, which means that the gas changes its condition without the absorption or rejection of heat either from or to an external body or source. It is common practice, in the study of cycles of refrigeration, to plot the compression line either along or parallel to a line of constant entropy. 
     In  FIG. 5 , line C-D denotes the compression process, in which the pressure and temperature of the vapor are increased from that in the evaporator  110  to that in the condenser  107 , with the assumption that there has been no pickup of heat in the suction line  111  between the evaporator  110  and the compressor  105 . For a condensing temperature of 100° F., a pressure gauge would read approximately 196 psig; but the chart is rated in absolute pressure and the atmospheric pressure of 14.7 is added to the psig, making it actually 210.61 psia. 
     Point Don the absolute pressure line is equivalent to the 100° F. condensing temperature; it is not on the saturated vapor line, it is to the right in the superheat area, at a junction of the 210.61 psia line, the line of constant entropy of 40° F., and the temperature line of approximately 128° F. A line drawn vertically downward from point D intersects the heat content line at 118.68 Btu/lb, which is h d , and the difference between h, and h d  is 10.54 Btu/lb—the heat of compression that has been added to the vapor. This amount of heat is the heat energy equivalent of the work done during the refrigeration compression cycle. This is the theoretical discharge temperature, assuming that saturated vapor enters the cycle; in actual operation, the discharge temperature may be 20° to 35° higher than that predicted theoretically. This can be checked in the system  100  by attaching a temperature sensor  1016  to the hot gas line  106 . 
     During the compression process, the vapor is heated by the action of its molecules being pushed or compressed closer together, commonly called heat of compression. 
     Line D-E denotes the amount of superheat that must be removed from the vapor before it can commence the condensation process. A line drawn vertically downward from point E to point h e  on the heat content line indicates the distance h d −h e , or heat amounting to 6.57 Btu/lb, since the heat content of 100° F. vapor is 112.11 Btu/lb. This superheat is usually removed in the hot gas discharge line or in the upper portion of the condenser  107 . During this process the temperature of the vapor is lowered to the condensing temperature. 
     Line E-A represents the condensation process that takes place in the condenser  107 . At point E the refrigerant is a saturated vapor at the condensing temperature of 100° F. and an absolute pressure of 210.61 psia; the same temperature and pressure prevail at point A, but the refrigerant is now in a liquid state. At any other point on line E-A the refrigerant is in the phase of a liquid vapor combination; the closer the point is to A, the greater the amount of the refrigerant that has condensed into its liquid stage. At point A, each pound of refrigerant is ready to go through the refrigerant cycle again as it is needed for heat removal from the load in the evaporator  110 . 
     Two factors that determine the coefficient of performance (COP) of a refrigerant are refrigerating effect and heat of compression. The equation (equation 1) may be written as: 
             COP   =     refrigerating_effect     heat_of   ⁢   _compression             
Substituting values, from the pressure-enthalpy diagram of the simple saturated cycle previously presented, the equation would be:
 
     
       
         
           
             COP 
             = 
             
               
                 
                   
                     h 
                     c 
                   
                   - 
                   
                     h 
                     a 
                   
                 
                 
                   
                     h 
                     d 
                   
                   - 
                   
                     h 
                     c 
                   
                 
               
               = 
               
                 
                   68.87 
                   10.54 
                 
                 = 
                 6.53 
               
             
           
         
       
     
     The COP is, therefore, a rate or a measure of the theoretical efficiency of a refrigeration cycle and is the energy that is absorbed in the evaporation process divided by the energy supplied to the gas during the compression process. As can be seen from Equation 1, the less energy expended in the compression process, the larger will be the COP of the refrigeration system. 
     The pressure-enthalpy diagrams in  FIGS. 4 and 5  show a comparison of two simple saturated cycles having different evaporating temperatures, to bring out various differences in other aspects of the cycle. In order that an approximate mathematical calculation comparison may be made, the cycles shown in  FIGS. 4 and 5  will have the same condensing temperature, but the evaporating temperature will be lowered 20° F. The values of A, B, C, D, and E from  FIG. 4  are compared to that in  FIG. 5  (with a 20° F. evaporator  110 ). The refrigerating effect, heat of compression, and the heat dissipated at the condenser  107  in each of the refrigeration cycles will be compared. The comparison will be based on data about the heat content or enthalpy line, rated in Btu/lb. 
     For the 20° F. evaporating temperature cycle shown in  FIG. 5 :
 
Net refrigerating effect( h   c   ′−h   a )=67.11 Btu/lb
 
Heat of compression( h   d   −h   c ′)=12.30 Btu/lb
 
     Comparing the data above with those of the cycle with the 40° F. evaporating temperature  FIG. 4 , shows that there is a decrease in the net refrigeration effect (NRE) of 2.6% and an increase in the heat of compression of 16.7%. There will be some increase in superheat, which should be removed either in the hot gas line  106  or the upper portion of the condenser  107 . This is the result of a lowering in the suction temperature, the condensing temperature remaining the same. 
     It follows that the weight of refrigerant to be circulated per ton of cooling, in a cycle with a 20° F. evaporating temperature and a 100° F. condensing temperature, is 2.98 lb/min/ton: 
     
       
         
           
             
               
                 
                   W 
                   = 
                     
                   ⁢ 
                   
                     
                       200 
                       ⁢ 
                       
                         ( 
                         
                           Btu 
                           ⁢ 
                           
                             / 
                           
                           ⁢ 
                           min 
                         
                         ) 
                       
                     
                     
                       NRE 
                       ⁡ 
                       
                         ( 
                         
                           Btu 
                           ⁢ 
                           
                             / 
                           
                           ⁢ 
                           lb 
                         
                         ) 
                       
                     
                   
                 
               
             
             
               
                 
                   = 
                     
                   ⁢ 
                   
                     
                       200 
                       ⁢ 
                       
                           
                       
                       ⁢ 
                       Btu 
                       ⁢ 
                       
                         / 
                       
                       ⁢ 
                       min 
                     
                     
                       67.11 
                       ⁢ 
                       
                           
                       
                       ⁢ 
                       Btu 
                       ⁢ 
                       
                         / 
                       
                       ⁢ 
                       lb 
                     
                   
                 
               
             
             
               
                 
                   = 
                     
                   ⁢ 
                   
                     2.98 
                     ⁢ 
                     
                         
                     
                     ⁢ 
                     lb 
                     ⁢ 
                     
                       / 
                     
                     ⁢ 
                     min 
                   
                 
               
             
           
         
       
     
     Circulating more refrigerant typically involves either a larger compressor  105 , or the same size of compressor  105  operating at a higher rpm. 
       FIG. 6  shows the original cycle with a 40° F. evaporating temperature, but the condensing temperature has been increased to 120° F. 
     Again taking the specific data from the heat content or enthalpy line, one now finds for the 120° F. condensing temperature cycle that h a ′=45.71, h c =108.14, h d ′=122.01, and h e ′=112.78. Thus, the net refrigerating effect (h c −h a ′)=62.43 Btu/lb, the heat of compression (h d ′−h c )=13.87 Btu/lb, and the condenser  107  superheat (h d ′−h e ′)=9.23 Btu/lb. 
     In comparison with the cycle having the 100° F. condensing temperature ( FIG. 4 ), the cycle can also be calculated by allowing the temperature of the condensing process to increase to 120° F. (as shown in  FIG. 6 ).  FIG. 6  shows a decrease in the NRE of 9.4%, an increase in heat of compression of 31.6%, and an increase of superheat to be removed either in the discharge line or in the upper portion of the condenser  107  of 40.5%. 
     With a 40° F. evaporating temperature and a 120° F. condensing temperature, the weight of refrigerant to be circulated will be 3.2 lb/min/ton. This indicates that approximately 10% more refrigerant must be circulated to do the same amount of work as when the condensing temperature was 100° F. 
     Both of these examples show that for the best efficiency of a system, the suction temperature should be as high as feasible, and the condensing temperature should be as low as feasible. Of course, there are limitations as to the extremes under which the system  100  may operate satisfactorily, and other means of increasing efficiency must then be considered. Economics of equipment (cost+operating performance) ultimately determine the feasibility range. 
     Referring to  FIG. 7 , after the condensing process has been completed and all of the refrigerant vapor at 120° F. is in the liquid state, if the liquid can be subcooled to point A on the 100° F. line (a difference of 20° F.), the NRE (h c −h a ) will be increased by 6.44 Btu/lb. This increase in the amount of heat absorbed in the evaporator  110  without an increase in the heat of compression will increase the COP of the cycle, since there is no increase in the energy input to the compressor  105 . 
     This subcooling can take place while the liquid is temporarily in storage in the condenser  107  or receiver, or some of the liquid&#39;s heat may be dissipated to the ambient temperature as it passes through the liquid pipe on its way to the metering device  109 . Subcooling can also take place in a commercial type water cooled system through the use of a liquid subcooler. 
     Normally, the suction vapor does not arrive at the compressor  105  in a saturated condition. As shown in  FIG. 8 , superheat is added to the vapor after the evaporating process has been completed, in the evaporator  110  and/or in the suction line  111 , as well as in the compressor  105 . If this superheat is added only in the evaporator  110 , it is doing some useful cooling; for it too is removing heat from the load or product, in addition to the heat that was removed during the evaporating process. But if the vapor is superheated in the suction line  111  located outside of the conditioned space, no useful cooling is accomplished; yet this is what takes place in many systems. 
     In the system  100 , the refrigerant pressure is relatively high in the condenser  107  and relatively low in the evaporator  110 . A pressure rise occurs across the compressor  105  and a pressure drop occurs across the metering device  109 . Thus, the compressor  105  and the metering device  109  maintain the pressure difference between the condenser  107  and the evaporator  110 . 
     Thus, a refrigeration system can be divided into the high side and low side portions. The high side contains the high pressure vapor and liquid refrigerant and is the part of the system that rejects heat. The low side contains the low pressure liquid vapor and refrigerant and is the side that absorbs heat. 
     Heat is always trying to reach a state of balance by flowing from a warmer object to a cooler object. Heat only flows in one direction, from warmer to cooler. Temperature difference (TD) is what allows heat to flow from one object to another. The greater the temperature difference the more rapid the heat flow. For the high side of a refrigeration unit to reject heat, its temperature must be above the ambient or surrounding temperature. For the evaporator  110  to absorb heat, its temperature must be below the surrounding ambient temperature. 
     Two factors that affect the quantity of heat transferred between two objects are the temperature difference and the mass of the two objects. The greater the temperature difference between the refrigerant coil (e.g., the condenser  107  or the evaporator  110 ) and the surrounding air, the more rapid will be the heat transfer. The larger the size of the refrigerant coil, the greater the mass of refrigerant, which also increases the rate of heat transfer. Engineers can either design coils to have high temperature differences or larger areas to increase the heat transfer rate. 
     To increase energy efficiency, systems are designed with larger coils because it is more efficient to have a lower temperature difference and a larger area to transfer heat. It takes less energy to produce a smaller pressure/temperature difference within a refrigeration system. Manufacturers of new high efficiency air conditioning systems use this principle. 
     The same principle can be applied to the evaporator  110  coils. The temperature differences between the evaporator input air  124  and the evaporator output air  125  are lower than they were on earlier systems. Older, lower efficiency air conditioning systems may have evaporative coils that operate at 35° F. output temperature, while newer higher efficiency evaporator  110  may operate in the 45° F. output range. Both evaporators  110  can pick up the same amount of heat provided that the higher temperature, higher efficiency coil has greater area and, therefore, more mass of refrigerant being exposed to the air stream to absorb heat. The higher evaporative coil temperature may produce less dehumidification. In humid climates, dehumidification can be an important part of the total air conditioning. 
     Correct equipment selection is important to ensure system operation and to obtain desired energy efficiencies. Previously, it was a common practice in many locations for installers to select an evaporator  110  of a different tonnage than the condenser unit  101  capacity. While this practice in the past may provide higher efficiencies, for most of today&#39;s more technically designed systems proper matching is usually achieved by using the manufacturer&#39;s specifications in order to provide proper operation. Mismatching systems can result in poor humidity control and higher operating costs. In addition to poor energy efficiency and lack of proper humidity control, the compressor  105  in a mismatched system may not receive adequate cooling from returning refrigerant vapor. As a result the compressor  105  temperature will be higher, and this can reduce the life of the compressor  105 . 
     As refrigerant vapor leaves the discharge side of the compressor  105 , it enters the condenser  107 . As this vapor travels through the condenser  107 , heat from the refrigerant dissipates to the surrounding air through the piping and fans. As heat is removed, the refrigerant begins to change state from vapor to liquid. As the mixture of liquid and vapor continues to flow through the condenser  107 , more heat is removed and eventually all, or virtually all, of the vapor has transformed into liquid. The liquid flows from the outlet of the condenser  107  through the liquid line  108  to the metering device  109 . 
     The high pressure, high temperature liquid refrigerant passes through the metering device  109  where its temperature and pressure change. As the pressure and temperature change, some of the liquid refrigerant boils off forming flash gas. As this mixture of refrigerant, liquid, and vapor flow through the evaporator  110 , heat is absorbed, and the remaining liquid refrigerant changes into a vapor. At the outlet of the evaporator  110  the vapor flows back through the suction line  111  to the compressor  105 . 
     The compressor  105  draws in this low pressure, low temperature vapor and converts it to a high temperature, high pressure vapor where the cycle begins again. 
     An ideally sized and functioning system  100  is one where the last bit of refrigerant vapor changes into a liquid at the end of the condenser  107  and where the last bit of liquid refrigerant changes into a vapor at the end of the evaporator  110 . However, because it is impossible to have a system operate at this ideal state, units are designed to have some additional cooling, called subcooling, of the liquid refrigerant to ensure that no vapor leaves the condenser  107 . Even a small amount of vapor leaving the condenser  107  can significantly reduce efficiency of the system  100 . 
     On the evaporator  110  side a small amount of additional temperature is added to the refrigerant vapor, called superheat, to ensure that no liquid refrigerant returns to the compressor  105 . Returning liquid refrigerant to the compressor  105  can damage the compressor  105 . 
     Systems that must operate under a broad range of temperature conditions will have difficulty maintaining the desired level of subcooling and superheat. There are two components that can be used in these systems to enhance the level of efficiency and safety in operation. They are the receiver and the accumulator. The receiver is placed in the liquid line  108  and holds a little extra refrigerant so the system has enough for high loads on hot days. The accumulator is placed in the suction line  111  and traps any the liquid refrigerant that would flow back to the compressor  105  on cool days with light loads. 
     A liquid receiver can be located at the end of the condenser  107  outlet to collect liquid refrigerant. The liquid receiver allows the liquid to flow into the receiver and any vapor collected in the receiver to flow back into the condenser  107  to be converted back into a liquid. The line connecting the receiver to the condenser  107  is called the condensate line and must be large enough in diameter to allow liquid to flow into the receiver and vapor to flow back into the condenser  107 . The condensate line must also have a slope toward the receiver to allow liquid refrigerant to freely flow from the condenser  107  into the receiver. The outlet side of the receiver is located at the bottom where the trapped liquid can flow out of the receiver into the liquid line. 
     Receivers should be sized so that all of the refrigerant charge can be stored in the receiver. Some refrigeration condensing units come with receivers built into the base of the condensing unit. 
     The accumulator is located at the end of the evaporator  110  and allows liquid refrigerant to be collected in the bottom of the accumulator and remain there as the vapor refrigerant is returned to the compressor  105 . The inlet side of the accumulator is connected to the evaporator  110  where any liquid refrigerant and vapor flow in. The outlet of the accumulator draws vapor through a U shaped tube or chamber. There is usually a small port at the bottom of the U shaped tube or chamber that allows liquid refrigerant and oil to be drawn into the suction line. Without this small port, refrigerant oil would collect in the accumulator and not return to the compressor  105 . The small port does allow some liquid refrigerant to enter the suction line. However, it is such a small amount of liquid refrigerant that it boils off rapidly, so there is little danger of liquid refrigerant flowing into the compressor  105 . 
     Accumulators are often found on heat pumps. During the changeover cycle, liquid refrigerant can flow back out of the outdoor coil. This liquid refrigerant could cause compressor  105  damage if it were not for the accumulator, which blocks its return. 
     The pressure-heat diagram of  FIG. 8  shows the cooling process in the evaporator  110 . Initially the high pressure liquid is usually subcooled 8-10° F. or more. When subcooled liquid from point A flows through the metering device  109 , its pressure drops to the pressure of the evaporator  110 . Approximately 20% of the liquid boils off to gas, cooling the remaining liquid-gas mixture. Its total heat (enthalpy) at point B is relatively unchanged from A. No external heat energy has been exchanged. From points B to C, the remainder of the liquid boils off, absorbing the heat flowing in from the evaporator  110  load (air, water, etc.). At point C, all of the liquid has evaporated and the refrigerant is vapor at the saturation temperature corresponding to the evaporator  110  pressure. 
     The subcooling increases cycle efficiency and can prevent flash gas due to pressure loss from components, pipe friction, or increase in height. 
     Many smaller refrigeration systems are designed to have the metering device  109  control the refrigerant flow so the evaporator  110  will heat the vapor beyond saturated conditions and ensure no liquid droplets will enter and possibly damage the compressor  105 . It is assumed here for the sake of simplicity there is no pressure drop through the evaporator  110 . In reality there are pressure drops which would slightly shift the evaporating and condensing processes from the constant pressure lines shown. 
     If the evaporator  110  does not have to superheat refrigerant vapor, it can produce more cooling capacity. On smaller systems the difference is relatively small and it is more important to protect the compressor  105 . On larger systems, an increase in evaporator performance can be important. A flooded evaporator  110  absorbs heat from points B to C. It can circulate more pounds of refrigerant (more cooling capacity) per square foot of heat transfer surface. 
     An undersized evaporator with less heat transfer surface will not handle the same heat load at the same temperature difference as a correctly sized evaporator. The new balance point will be reached with a lower suction pressure and temperature. The load will be reduced and the discharge pressure and temperature will also be reduced. An undersized evaporator and a reduced heat load both have similar effects on the refrigerant cycle because they both are removing less heat from the refrigerant. 
     As the ambient temperature increases, the load on the evaporator  110  increases. When the load on the evaporator  110  increases, the pressures increase. The operating points shift up and to the right on the pressure-heat curve. As the load on the evaporator  110  decreases, the pressures decrease. The operating points on the pressure-heat curve shift down. Thus, knowledge of the ambient temperature is useful in determining whether the system  100  is operating efficiency. 
       FIG. 9A  is a block diagram of a monitoring system  900  for monitoring the operation of the refrigerant-cycle system. In  FIG. 9A , one or more condenser unit sensors  901  measure operating characteristics of the elements of the condenser unit  101 , one or more evaporator unit sensors  902  measure operating characteristics of the evaporator unit  102 , and one or more ambient sensors  903  measure ambient conditions. Sensor data from the condenser unit sensors  901 , evaporator unit sensors  902 , and ambient sensors  903  are provided to a processing system  904 . The processing system  904  uses the sensor data to calculate system efficiency, identify potential performance problems, calculate energy usage, etc. In one embodiment, the processing system  904  calculates energy usage and energy costs due to inefficient operation. In one embodiment, the processing system  904  schedules filter maintenance according to elapsed time and/or filter usage. In one embodiment, the processing system  904  identifies potential performance problems (e.g., low airflow, insufficient or unbalanced load, excessive load, low ambient temperature, high ambient temperature, refrigerant undercharge, refrigerant overcharge, liquid line restriction, suction line restriction, hot gas line restriction, inefficient compressor, etc.). In one embodiment, the processing system  904  provides plots or charts of energy usage and costs. In one embodiment, the processing system  904  of the system  900  provides plots or charts of the additional energy costs due to inefficient operation of the refrigerant-cycle system. In one embodiment, a thermostat  952  is provided to the processing system  904 . In one embodiment, the processing system  904  and thermostat  952  are combined. 
       FIG. 9B  is a block diagram of the system  900  wherein operating data from the refrigerant-cycle system is provided to a remote monitoring system  950 , such as, for example, a power company or monitoring center. In one embodiment, the system  900  provides operating data related to the operating efficiency of the refrigerant-cycle system to the remote monitoring system  950 . In one embodiment, the remote monitoring system  950  provides operating efficiency data to an electric power company or governmental agency. 
     Data can be transmitted from the system  900  to a remote monitoring system  950  by using data transmission over power lines as shown in  FIG. 9B  and/or by using data transmission over a data network (e.g., the Internet, a wireless network, a cable modem network, a telephone network, etc.) as shown in  FIG. 9C  and also as shown in discussed in connection with  FIGS. 9F-9H . 
       FIG. 9D  is a block diagram of a monitoring system for monitoring the operation of the refrigerant-cycle system, where data regarding operation of the system is provided to a thermostat  952  and/or to a computer system  948  such as, for example, a site monitoring computer, a maintenance computer, a personal digital assistant, a personal computer, etc. 
       FIG. 9E  is a block diagram of a monitoring system for monitoring the operation of the refrigerant-cycle system wherein an electronically-controlled metering device  960  is provided to allow control of the system in an energy-efficient matter. 
       FIG. 9F  is a block diagram of a thermostat control and monitoring system having a data interface device  955  provided to the thermostat  952 . The thermostat  952  typically communicates with an evaporator unit controller  953  using relatively low-voltage control wiring. The evaporator unit controller  953  typically provides relays and other control circuits for the air handler fan, and other systems in the evaporator unit  102 . The control wiring is also provided to a condenser unit controller  954  in the condenser unit  101 . The condenser unit controller  954  provides relays and other control circuits for the compressor  105 , the condenser fan, etc. The data interface device  955  is provided to the low-voltage control wiring to allow the thermostat  952  to receive control signals from the remote monitoring system  950 . 
       FIG. 9G  is a block diagram of a thermostat control and monitoring system wherein a data interface device  956  is provided to the condenser unit controller  954 . The data interface device  956  allows the remote monitoring system  950  to communicate with the condenser unit  101 . In one embodiment, the data interface device  956  allows the remote monitoring system  950  to read sensor data from the condenser unit  101 . In one embodiment, the data interface device  956  allows the remote monitoring system  950  to turn off the condenser unit  101 . In one embodiment, the data interface device  956  allows the remote monitoring system  950  to switch the compressor  105  to a lower-speed mode. In one embodiment, the data interface device  956  allows the remote monitoring system  950  to switch the condenser unit  101  to a power conservation mode. 
       FIG. 9H  is a block diagram of a thermostat control and monitoring system wherein a data interface device  957  is provided to the evaporator unit controller  953 . 
     In one embodiment, the data interface devices  955 - 957  are configured as power line modems (e.g., using Broadband over Power Line (BPL), or other power line networking technology). In one embodiment, the data interface devices  955 - 957  are configured as wireless modems for communication using wireless transmission. In one embodiment, the data interface devices  955 - 957  are configured as telephone modems, cable modems, Ethernet modems, or the like, to communicate using a wired network. 
     In one embodiment, the system  900  provides sensor data from the condenser unit sensors  901  and/or the evaporator unit sensors  902  to the remote monitoring system  950 . In one embodiment, the system  900  uses data from the condenser unit sensors  901  and/or the evaporator unit sensors  902  to compute an efficiency factor for the refrigerant-cycle system and the system  900  provides the efficiency factor to the remote monitoring system  950 . In one embodiment, the system  900  provides power usage data (e.g., amount of power used) by the refrigerant-cycle system and the system  900  provides the efficiency factor to the remote monitoring system  950 . In one embodiment, the system  900  provides an identification code (ID) with the data transmitted to the remote monitoring system  950  to identify the system  900 . 
     In one embodiment, the remote monitoring system  950  is provided with data regarding a maximum expected efficiency for the refrigerant-cycle system (e.g., based on the manufacture and design characteristics of the refrigerant-cycle system) such that the remote monitoring system  950  can ascertain the relative efficiency (that is, how the refrigerant-cycle system is operating with respect to its expected operating efficiency). In one embodiment, the remote monitoring system  950  provides efficiency data to the power company or to a government agency so electric rates can be charged according to the system efficiency. In one embodiment, the homeowner (or building owner) is charged a higher electrical rate for electrical power provided to a refrigerant-cycle system that is operating at a relatively low absolute efficiency. In one embodiment, the homeowner (or building owner) is charged a higher electrical rate for electrical power provided to a refrigerant-cycle system that is operating at a relatively low relative efficiency. In one embodiment, the homeowner (or building owner) is charged an electrical rate according to a combination of the relative and absolute efficiency of the refrigerant-cycle system. In one embodiment, the data provided to the remote monitoring system  950  is used to provide notice to the homeowner (or building owner) that the refrigerant-cycle system is operating at a poor efficiency. In one embodiment, the data provided to the remote monitoring system  950  is used to provide notice to the homeowner (or building owner) that the refrigerant-cycle system is operating at a poor efficiency, and that the system must be serviced. In one embodiment, the owner is given a warning that service is needed. If the unit is not serviced (or if efficiency does not improve) after a period of time, the remote monitoring system  950  can remotely shut off the refrigerant-cycle system by sending commands to one or more of the data interface devices  955 - 957 . 
     In one embodiment, the homeowner (or building owner) is charged a higher electrical rate for electrical power provided to a refrigerant-cycle system that is operating at a relatively low efficiency during a specified period of time, such as, for example, when the power system is highly loaded, during peak afternoon cooling periods, during heat waves, during rolling blackouts, etc. In one embodiment, the homeowner (or building owner) is charged a higher electrical rate (a premium rate) for electrical power provided to a refrigerant-cycle system during a specified period of time, such as, for example, when the power system is highly loaded, during peak afternoon cooling periods, during heat waves, during rolling blackouts, etc. In one embodiment, the homeowner (or building owner) can program the system  900  to receive messages from the power company indicating that premium rates are being charged. In one embodiment, the homeowner (or building owner) can program the system  900  to shut down during premium rate periods. In one embodiment, the homeowner (or building owner) can avoid paying premium rates by allowing the power company to remotely control operation of the refrigerant-cycle system during premium rate times. In one embodiment, the homeowner (or building owner) is only allowed to run the refrigerant-cycle system during premium rate periods if the system is operating above a prescribed efficiency. 
     In one embodiment, the system  900  monitors the amount of time that the refrigerant-cycle system has been running (e.g., the amount of runtime during the last day, week, etc.). In one embodiment, the remote monitoring system  950  can query the system  900  to obtain data regarding the operating of the refrigerant-cycle system and one or more of the data interface devices  955 - 957  will receive the query and send the requested data to the remote monitoring system  950 . The query data may be, for example, the efficiency rating of the refrigerant-cycle system (e.g., the SEER, EER, etc.), the current operating efficiency of the refrigerant-cycle system, the runtime of the system during a specified time period, etc. The remote monitoring system  950  operator (e.g., the power company or power transmission company) can use the query data to make load balancing decisions. Thus, for example the decision regarding whether to instruct the refrigerant-cycle system to shut down or go into a low power mode can be based on the system efficiency (specified efficiency, absolute efficiency, and/or relative efficiency), the amount of time the system has been running, the home or building owner&#39;s willingness to pay premium rates during load shedding periods, etc. Thus, for example a homeowner who has a low-efficiency system that is heavily used, or who has indicated an unwillingness to pay premium rates, would have his/her refrigerant-cycle system shut off by the remote monitoring system  950  before that of a homeowner who has installed a high-efficiency system that is used relatively little, and who had indicated a willingness to pay premium rates. In one embodiment, in making the decision to shut off the system  900 , the remote monitoring system  950  would take into account the efficiency of the system  900 , the amount the system  900  is being used, and the owner&#39;s willingness to pay premium rates. In one embodiment, higher-efficiency systems are preferred over lower-efficiency systems (that is, higher-efficiency systems are less likely to be shut off during a power emergency), and lightly-used systems are preferred over heavily-used systems. 
     In one embodiment, the system  900  sends data regarding the set temperature of the thermostat  952  to the remote monitoring system  950 . In one embodiment, the electricity rate charged to the homeowner (or building owner) is calculated according to a set point of the thermostat  952  such that a lower set point results in a higher rate charge per kilowatt-hour. In one embodiment, the electricity rate charged to the homeowner (or building owner) is calculated according to the set point of the thermostat  952  and the relative efficiency of the refrigerant-cycle system such that a lower set point and/or lower efficiency results in a higher rate charge per kilowatt-hour. In one embodiment, the electricity rate charged to the homeowner (or building owner) is calculated according to the set point of the thermostat  952  and the absolute efficiency of the refrigerant-cycle system such that a lower set point and/or lower efficiency results in a higher rate charge per kilowatt-hour. In one embodiment, the electricity rate charged to the homeowner (or building owner) is calculated according to the set point of the thermostat  952 , the relative efficiency of the refrigerant-cycle system, and the absolute efficiency of the refrigerant-cycle system according to a formula whereby a lower set point and/or lower efficiency results in a higher rate charge per kilowatt-hour. 
     In one embodiment, the remote monitoring system  950  can send instructions to the system  900  to shut down if the refrigerant-cycle system is operating at a low efficiency. In one embodiment, the remote monitoring system  950  can send instructions to the system  900  to change the setting of the thermostat  952  (e.g., raise the set temperature of the thermostat  952 ) in response to low efficiency of the refrigerant-cycle system and/or to avoid a blackout. In one embodiment the remote monitoring system  950  can send instructions to the condenser unit  101  to switch the compressor  105  to a low-speed mode to conserve power. 
     In one embodiment, the remote monitoring system  950  knows the identification codes or addresses of the data interface devices  955 - 957  and correlates the identification codes with a database to determine whether the refrigerant-cycle system is serving a relatively high priority client such as, for example, a hospital, the home of an elderly or invalid person, etc. In such circumstances, the remote monitoring system  950  can provide relatively less cutback in cooling provided by the refrigerant-cycle system. 
     In one embodiment, the system  900  communicates with the remote monitoring system  950  to provide load shedding. Thus, for example, the remote monitoring system  950  (e.g., a power company) can communicate with the data interface device  956  and/or the data interface device  957  to turn off the refrigerant-cycle system. The remote monitoring system  950  can thus rotate the on and off times of air conditioners across a region to reduce the power load without implementing rolling blackouts. In one embodiment, the data interface device  956  is configured as a retrofit device that can be installed in a condenser unit to provide remote shutdown. In one embodiment, the data interface device  956  is configured as a retrofit device that can be installed in a condenser unit to remotely switch the condenser-unit to a low power (e.g., energy conservation) mode. In one embodiment, the data interface device  957  is configured as a retrofit device that can be installed in an evaporator unit to provide remote shutdown or to remotely switch the system to a lower power mode. In one embodiment, the remote monitoring system  950  sends separate shutdown and restart commands to one or more of the data interface devices  955 - 957 . In one embodiment, the remote monitoring system  950  sends commands to the data interface devices  955 - 957  to shutdown for a specified period of time (e.g., 10 min, 30 min, 1 hour, etc.) after which the system automatically restarts. 
     In one embodiment, the system  900  communicates with the remote monitoring system  950  to control the temperature set point of the thermostat  952  to prevent blackouts or brownouts without regard to efficiency of the refrigerant-cycle system. When brownout or potential blackout conditions occur, the remote monitoring system  950  can override the homeowner&#39;s thermostat setting to cause the temperature set point on the thermostat  952  to change (e.g., increase) in order to reduce power usage. In most residential installations, low-voltage control wiring is provided between the thermostat  952  and the evaporator unit  102  and condenser unit  101 . In most residential (and many industrial) applications the thermostat  952  receives electrical power via the low-voltage control wiring from a step-down transformer provided with the evaporator unit  102 . 
     In one embodiment, the data interface device  955  is provided in connection with the power meter  949 , and the data interface device  955  communicates with the thermostat  952  using wireless communications. 
     In a typical refrigeration or air conditioning system, the condenser unit  101  is placed outside the area being cooled and the evaporator unit  102  is placed inside the area being cooled. The nature of outside and inside depend on the particular installation. For example, in an air conditioning or HVAC system, the condenser unit  101  is typically placed outside the building, and the evaporator unit  102  is typically placed inside the building. In a refrigerator or freezer, the condenser unit  101  is placed outside the refrigerator and the evaporator unit  102  is placed inside the refrigerator. In any case, the waste heat from the condenser should be dumped outside (e.g., away from) the area being cooled. 
     When the system  900  is installed, the system  900  is programmed by specifying the type of refrigerant used and the characteristics of the condenser  107 , the compressor  105 , and the evaporator unit  102 . In one embodiment, the system  900  is also programmed by specifying the size of the air handler system. In one embodiment, the system  900  is also programmed by specifying the expected (e.g., design) efficiency of the system  100 . 
     The remote monitoring system  950  can do a better job of monitoring efficiency than published performance ratings such as the Energy Efficiency Ratio (EER) and SEER. The EER is determined by dividing the published steady state capacity by the published steady state power input at 80° F. DB/67° F. WB indoor and 95° F. DB outdoor. This is objective yet unrealistic with respect to system “real world” operating conditions. The published SEER rating of a system is determined by multiplying the steady state EER measured at conditions of 82° F. outdoor temperature, 80° F. DB/67° F. WB indoor entering air temperature by the (run time) Part Load Factor (PLF) of the system. A major factor not considered in SEER calculations is the actual part loading factor of the indoor evaporator cooling coil, which reduces the unit&#39;s listed BTUH capacity and SEER efficiency level. Many older air handlers and duct systems do not deliver the published BTUH and SEER Ratings. This is primarily due to inadequate air flow through the evaporator  110 , a dirty evaporator  110 , and/or dirty blower wheels. Also, improper location of supply diffusers and return air registers can result in inefficient floor level recirculation of the cold conditioned air, resulting in lack of heat loading of the evaporator  110 . 
     By monitoring the system  100  under actual load conditions, and by measuring the relevant ambient temperature and humidity, the system  900  can calculate the actual efficiency of the system  100  in operation. 
       FIG. 10  shows a monitoring system  1000  for monitoring the operation of the system  100 . The system  1000  shown in  FIG. 10  is one example of an embodiment of the system  900  shown in  FIGS. 9A-E . In the system  1000 , a condenser unit sender  1002  monitors operation of the condenser unit  101  through one or more sensors, an evaporator sender unit  1003  monitors operation of the evaporator unit  102  through one or more sensors. The condenser unit sender  1002  and the evaporator sender unit  1003  communicate with the thermostat  1001  to provide data to the building owner. For purposes of explanation, and not by way of limitation, in  FIG. 10  the processing system  904  and thermostat  952  from  FIGS. 9A-E  are shown as a single thermostat-processor. One of ordinary skill in the art will recognize that the processor functions can be separated from the thermostat. 
     In one embodiment, a building interior temperature sensor  1009  is provided to the thermostat  1001 . In one embodiment, a building interior humidity sensor  1010  is provided to the thermostat  1001 . In one embodiment, the thermostat  1001  includes a display  1008  for displaying system status and efficiency. In one embodiment, the thermostat  1001  includes a keypad  1050  and/or indicator lights (e.g., LEDs)  1051 . A power sensor  1011  to sense electrical power consumed by the compressor  105  is provided to the condenser unit sender  1002 . In one embodiment, a power sensor  1017  to sense electrical power consumed by the condenser fan  122  is provided to the condenser unit sender  1002 . The air  125  from the evaporator  110  flows in the ductwork  1080 . 
     In one embodiment, a temperature sensor  1012 , configured to measure the temperature of the refrigerant in the suction line  111  near the compressor  105 , is provided to the condenser unit sender  1002 . In one embodiment, a temperature sensor  1016 , configured to measure the temperature of the refrigerant in the hot gas line  106 , is provided to the condenser unit sender  1002 . In one embodiment, a temperature sensor  1014 , configured to measure the temperature of the refrigerant in the liquid line  108  near the condenser  107 , is provided to the condenser unit sender  1002 . 
     Contaminants in the refrigerant lines  111 ,  106 ,  108 , etc. can reduce the efficiency of the refrigerant-cycle system and can reduce the life of the compressor  105  or other system components. In one embodiment, one or more contaminant sensors  1034 , configured to sense contaminants in the refrigerant (e.g., water, oxygen, nitrogen, air, improper oil, etc.) are provided in at least one of the refrigerant lines and provided to the condenser unit sender  1002  (or, optionally, to the evaporator unit sender  1003 ). In one embodiment, a contaminant sensor  1034  senses refrigerant fluid or droplets at the input to the compressor  105 , which can cause damage to the compressor  105 . In one embodiment, a contaminant sensor  1060  is provided in the liquid line  108  to sense bubbles in the refrigerant. Bubbles in the liquid line  108  may indicate low refrigerant levels, an undersized condenser  107 , insufficient cooling of the condenser  107 , etc. In one embodiment, the sensor  1034  senses water or water vapor in the refrigerant lines. In one embodiment, the sensor  1034  senses acid in the refrigerant lines. In one embodiment, the sensor  1034  senses air or other gasses (e.g., oxygen, nitrogen, carbon dioxide, chlorine, etc.). 
     In one embodiment, a pressure sensor  1013 , configured to measure pressure in the suction line  111 , is provided to the condenser unit sender  1002 . In one embodiment, a pressure sensor  1015 , configured to measure pressure in the liquid line  108 , is provided to the condenser unit sender  1002 . In one embodiment, a pressure sensor (not shown), configured to measure pressure in the hot gas line  106 , is provided to the condenser unit sender  1002 . In one embodiment, the pressure sensor  1013  and the pressure sensor  1015  are connected to the system  100 , by attaching the pressure sensors  1013  and  1015  to the service valves  120  and  121 , respectively. Attaching the pressure sensors to the service valves  120  and  121  is a convenient way to access refrigerant pressure in a retrofit installation without having to open the pressurized refrigerant system. 
     In one embodiment, a flow sensor  1031 , configured to measure flow in the suction line  111 , is provided to the condenser unit sender  1002 . In one embodiment, a flow sensor  1030 , configured to measure flow in the liquid line  108 , is provided to the condenser unit sender  1002 . In one embodiment, a flow sensor (not shown), configured to measure flow in the hot gas line  106 , is provided to the condenser unit sender  1002 . In one embodiment, the flow sensors are ultrasonic sensors that can be attached to the refrigerant lines without opening the pressurized refrigerant system. 
     In one embodiment, a temperature sensor  1028  configured to measure ambient temperature is provided to the condenser unit sender  1002 . In one embodiment, a humidity sensor  1029  configured to measure ambient humidity is provided to the condenser unit sender  1002 . 
     In one embodiment, a temperature sensor  1020 , configured to measure the temperature of the refrigerant in the liquid line  108  near the evaporator  110  is provided to the evaporator sender unit  1003 . In one embodiment, a temperature sensor  1021 , configured to measure the temperature of the refrigerant in the suction line  111  near the evaporator  110  is provided to the evaporator sender unit  1003 . 
     In one embodiment, a temperature sensor  1026 , configured to measure the temperature of air  124  flowing into the evaporator  110  is provided to the evaporator sender unit  1003 . 
     In one embodiment, a temperature sensor  1026 , configured to measure the temperature of air  125  flowing out of the evaporator  110  is provided to the evaporator sender unit  1003 . In one embodiment, an airflow sensor  1023 , configured to measure the airflow of air  125  flowing out of the evaporator  110  is provided to the evaporator sender unit  1003 . In one embodiment, a humidity sensor  1024 , configured to measure the humidity of air  125  flowing out of the evaporator  110  is provided to the evaporator sender unit  1003 . In one embodiment, a differential pressure sensor  1025 , configured to measure a pressure drop across the evaporator  110 , is provided to the evaporator sender unit  1003 . 
     In one embodiment, the temperature sensors are attached to the refrigerant lines (e.g., the lines  106 ,  108 ,  111 , in order to measure the temperature of the refrigerant circulating inside the lines. In one embodiment, the temperature sensors  1012  and/or  1016  are provided inside the compressor  105 . In one embodiment, the temperature sensors are provided inside one or more of the refrigerant lines. 
     A tachometer  1033  senses rotational speed of the fan blades in the fan  123 . The tachometer  1033  is provided to the evaporator unit sender  1003 . A tachometer  1032  senses rotational speed of the fan blades in the condenser fan  122 . The tachometer  1032  is provided to the condenser unit sender  1002 . 
     In one embodiment, a power sensor  1027 , configured to measure electrical power consumed by the fan  123  is provided to the evaporator sender unit  1003 . 
     In one embodiment, the evaporator sender unit  1003  communicates sensor data to the condenser unit sender  1002  through wireless transmission. In one embodiment, the evaporator sender unit  1003  communicates sensor data to the condenser unit sender  1002  through existing HVAC wiring. In one embodiment, the evaporator sender unit  1003  communicates sensor data to the condenser unit sender  1002  through existing HVAC wiring by modulating sensor data onto a carrier that is transmitted using the existing HVAC wiring. 
     Each of the sensors shown in  FIG. 10  (e.g., the sensors  1010 - 1034  etc.) are optional. The system  1000  can be configured with a subset of the illustrated sensors in order to reduce cost at the expense of monitoring system capability. Thus, for example, the contaminant sensors  1034  can be eliminated, but ability of the system  1000  to detect the contaminants sensed by the sensor  1034  will be compromised or lost. 
     The pressure sensors  1013  and  1015  measure suction and discharge pressures, respectively, at the compressor  105 . The temperature sensors  1026  and  1022  measure evaporator  110  return air temperature and supply air temperature, respectively. The temperature sensors  1018  and  1019  measure input air and discharge air, respectively, at the condenser  107 . 
     The power sensors  1011 ,  1017 , and  1027  are configured to measure electric power. In one embodiment, one or more of the power sensors measure voltage provided to a load and power is computed by using a specified impedance for the load. In one embodiment, one or more of the power sensors measure current provided to a load and power is computed by using a specified impedance for the load. In one embodiment, one or more of the power sensors measure voltage and current provided to a load and power is computed by using a specified power factor for the load. In one embodiment, the power sensors measure voltage, current, and the phase relationship between the voltage and the current. 
     The temperature sensors  1012  and/or  1021  measure the temperature of the refrigerant at the suction line  111 . By measuring the suction line  111  temperature, the superheat can be determined. The suction pressure has been measured by the pressure sensor  1013 , so the evaporating temperature can be read from a pressure-temperature chart. The superheat is the difference between the suction line  111  temperature and the evaporating temperature. 
     The temperature sensors  1014  and/or  1020  measure the temperature of the refrigerant in the liquid line  108 . By measuring the liquid line  108  temperature, the subcooling can be determined. The discharge pressure is measured by the pressure sensor  1015 , and thus the condensing temperature can be read from the pressure-temperature chart. The subcooling is the difference between the liquid line  108  temperature and the condensing temperature. 
     In one embodiment, the system  1000  calculates efficiency by measuring the work (cooling) done by the refrigerant-cycle system and dividing by the power consumed by the system. In one embodiment, the system  1000  monitors the system  100  for abnormal operation. Thus, for example, in one embodiment, the system  1000  measures the refrigerant temperature drop across the condenser  107  using the temperature sensors  1016  and  1014  to be used in calculating the heat removed by the condenser  107 . The system  1000  measures the refrigerant temperature drop across the evaporator  110  to be used in calculating the heat absorbed by the evaporator  110 . 
     The monitoring system  1000  is typically used to monitor the operation of the system  100  that was originally checked out and put into proper operation condition. Mechanical problems in an air conditioning system are generally classified in two categories: air side problems and refrigeration side problems. 
     The primary problem that can occur in the air category is a reduction in airflow. Air handling systems do not suddenly increase in capacity, that is, increase the amount of air across the coil. On the other hand, the refrigeration system does not suddenly increase in heat transfer ability. The system  1000  uses the temperature sensors  1026  and  1022  to measure the temperature drop of the air through the evaporator  110 . After measuring the return air and supply air temperatures and subtracting to get the temperature drop, the system  1000  checks to see whether the temperature difference is higher or lower than it should be. 
       FIG. 11  shows the temperature drop in the air through the evaporator  110  as a function of humidity. In one embodiment, the humidity sensors  1024  and/or  1010  are used to measure building humidity, and/or the humidity sensor  1029  is used to measure ambient humidity. The humidity readings are used to correct temperature readings for wet bulb temperature according to relative humidity. 
     In one embodiment, a comparison of the desired (or expected) temperature drop across the evaporator  110  with the measured actual temperature drop is used to help classify potential air problems from refrigerant cycle problems. If the actual temperature drop is less than the required temperature drop, then the airflow has likely been reduced. Reduced airflow can be caused by dirty air filters or evaporator  110 , problems with the fan  123 , and/or unusual restrictions in the duct system. 
     Air filters of the throwaway type are typically replaced at least twice each year, at the beginning of both the cooling and heating seasons. In one embodiment, the thermostat  1001  allows the owner to indicate when a new air filter is installed. The thermostat  1001  keeps track of the time the filter has been in use, and provides a reminder to the owner when the filter should be replaced. In one embodiment, the thermostat  1001  uses actual elapsed clock time to determine filter usage. 
     In one embodiment, the thermostat  1001  calculates filter usage according to the amount of time the air handler has been blowing air through the filter. Thus, for example, in moderate climates or seasons where the air handler system is not used continuously, the thermostat  1001  will wait a longer period of actual time before indicating that filter replacement is warranted. In some areas of higher use or where dust is high, the filter will generally have to be replaced relatively more often. In one embodiment, the thermostat  1001  uses a weighting factor to combine running time with idle time to determine filter usage. Thus, for example, in determining filter usage, hours when the air handler is blowing air thorough the filter are weighted relatively more heavily than hours where the air handler system is idle. In one embodiment, the owner can program the thermostat  1001  to indicate that filter replacement is needed after a specified number of hours or days (e.g., as actual days, as running days, or as a combination thereof). 
     In one embodiment, the thermostat  1001  is configured to receive information from an information source regarding daily atmospheric dust conditions and to use such information in calculating filter usage. Thus, in one embodiment, when calculating filter use, the thermostat  1001  weighs days of relatively high atmospheric dust relatively more heavily than days of relatively low atmospheric dust. In one embodiment, the information source for atmospheric dust information includes a data network, such as, for example, the Internet, a pager network, a local area network, etc. 
     In one embodiment, the thermostat  1001  collects data for calculating filter usage and passes such data to a computer monitoring system. 
     In commercial and industrial applications, a regular schedule of maintenance is generally used. In one embodiment, sensors are provided in connection with the air filter, as described below in connection with  FIG. 15 . 
     In one embodiment, power measured by the power sensor  1027  is used to help diagnose and detect problems with the fan  123  and/or the air handler system. If the fan  123  is drawing too much or too little current, or if the fan  123  is showing a low power factor, then possible problems with the blower and/or air handler system are indicated. 
     Placing furniture or carpeting over return air grilles reduces the air available for the blower to handle. Shutting off the air to unused areas will reduce the air over the evaporator  110 . Covering a return air grille to reduce the noise from the centrally located furnace or air handler may reduce the objectionable noise, but it also drastically affects the operation of the system by reducing the air quantity. The collapse of the return air duct system will affect the entire duct system performance. Air leaks in the return duct will raise the return air temperature and reduce the temperature drop across the coil. 
     The airflow sensor  1023  can be used to measure air flow through the ducts. In one embodiment, the airflow sensor  1023  is a hot wire (or hot film) mass flow sensor. In one embodiment, the differential pressure sensor  1025  is used to measure airflow through the evaporator  110 . In one embodiment, the differential pressure sensor  1025  is used to measure drop across the evaporator  110 . In one embodiment, the pressure drop across the evaporator  110  is used to estimate when the evaporator  110  is restricting airflow (e.g., due to damage, dirt, hair, dust, etc.). In one embodiment, the differential pressure sensor  1025  is used to measure drop across an air filter to estimate when the filter is restricting airflow (e.g., due to age, damage, dirt, hair, dust, etc.). In one embodiment, the indicator lights  1051  are used to indicate that the filter needs to be changed. In one embodiment, the indicator lights  1051  are used to indicate that the evaporator  110  needs to be cleaned. 
     In one embodiment, the airflow sensor  1023  is used to measure airflow into the ductwork  1080 . In one embodiment, the indicator lights  1051  are used to indicate that the airflow into the ductwork  1080  is restricted (e.g., due to dirt, furniture or carpets placed in front of vents, closed vents, dirty evaporator, dirty fan blades, etc.). 
     In one embodiment, a dust sensor is provided in the air stream of the evaporator  110 . In one embodiment, the dust sensor includes a light source (optical and/or infrared) and a light sensor. The dust sensor measures light transmission between the source and the light sensor. The buildup of dust will cause the light to be attenuated. The sensor detects the presence of dust buildup at the evaporator  110  by measuring light attenuation between the light source and the light sensor. When the attenuation exceeds a desired value, the system  1000  indicates that cleaning of the air flow system is needed (e.g., the fan  123 , the ductwork  1080 , and/or the evaporator  110 , etc.). 
     In one embodiment, the power sensor  1027  is used to measure power provided to the blower motor in the fan  123 . If the fan  123  is drawing too much power or too little power, then potential airflow problems are indicated (e.g., blocked or closed vents, dirty fan blades, dirty evaporator, dirty filter, broken fan belt, slipping fan belt, etc.). 
     If the temperature drop across the evaporator  110  is less than desired, then the heat removal capacity of the system has been reduced. Such problems can generally be divided into two categories: refrigerant quantity, and refrigerant flow rate. If the system  100  has the correct amount of refrigerant charge and refrigerant is flowing at the desired rate (e.g., as measured by the flow sensors  1031  and/or  1030 ), the system  100  should work efficiently and deliver rated capacity. Problems with refrigerant quantity or flow rate typically affect the temperatures and pressures that occur in the refrigerant-cycle system when the correct amount of air is supplied through the evaporator  110 . If the system is empty of refrigerant, a leak has occurred, and it must be found and repaired. If the system will not operate at all, it is probably an electrical problem that must be found and corrected. 
     If the system  100  will start and run but does not produce satisfactory cooling, then the amount of heat picked up in the evaporator  110  plus the amount of motor heat added and the total rejected from the condenser  107  is not the total heat quantity the unit is designed to handle. To diagnose the problem, the information listed in Table 1 is used. These results compared to normal operating results will generally identify the problem: (1) Evaporator  110  operating temperature; (2) Condensing unit condensing temperature; and/or (3) Refrigerant subcooling. 
     These items can be modified according to the expected energy efficiency ratio (EER) of the unit. The amount of evaporation and condensing surface designed into the unit are the main factors in the efficiency rating. A larger condensing surface results in a lower condensing temperature and a higher EER. A larger evaporating surface results in a higher suction pressure and a higher EER. The energy efficiency ratio for the conditions is calculated by dividing the net capacity of the unit in Btu/hr by the watts input. 
     
       
         
               
               
               
               
               
               
             
               
               
               
               
               
               
               
             
           
               
                 TABLE 1 
               
               
                   
               
               
                   
                   
                   
                   
                 Condenser 
                   
               
               
                   
                 Suction 
                 Evaporator 
                 Hot Gas 
                 Liquid 
                 Compressor 
               
               
                   
                 Pressure 
                 Superheat 
                 Pressure 
                 Subcooling 
                 Current 
               
               
                 Probable Cause 
                 (psig) 
                 (° F.) 
                 (psig) 
                 (° F.) 
                 (A) 
               
               
                   
               
             
             
               
                   
               
             
          
           
               
                 1. 
                 Insufficient or 
                 Low 
                 Low 
                 Low 
                 Normal 
                 Low 
               
               
                   
                 unbalanced load 
               
               
                 2. 
                 Excessive load 
                 High 
                 High 
                 High 
                 Normal 
                 High 
               
               
                 3. 
                 Low ambient 
                 Low 
                 High 
                 Low 
                 Normal 
                 Low 
               
               
                   
                 temperature 
               
               
                 4. 
                 High ambient 
                 High 
                 High 
                 High 
                 Normal 
                 High 
               
               
                   
                 temperature 
               
               
                 5. 
                 Refrigerant 
                 Low 
                 High 
                 Low 
                 Low 
                 Low 
               
               
                   
                 undercharge 
               
               
                 6. 
                 Refrigerant 
                 High 
                 Low 
                 High 
                 High 
                 High 
               
               
                   
                 overcharge 
               
               
                 7. 
                 Liquid line 
                 Low 
                 High 
                 Low 
                 High 
                 Low 
               
               
                   
                 restriction 
               
               
                 8. 
                 Plugged capillary 
                 Low 
                 High 
                 High 
                 High 
                 Low 
               
               
                   
                 tube 
               
               
                 9. 
                 Suction line 
                 Low 
                 High 
                 Low 
                 Normal 
                 Low 
               
               
                   
                 restriction 
               
               
                 10. 
                 Hot gas line 
                 High 
                 High 
                 High 
                 Normal 
                 High 
               
               
                   
                 restriction 
               
               
                 11. 
                 Inefficient 
                 High 
                 High 
                 Low 
                 Low 
                 Low 
               
               
                   
                 compressor 
               
               
                   
               
             
          
         
       
     
     Normal evaporator  110  operating temperatures can be found by subtracting the design coil split from the average air temperature going through the evaporator  110 . The coil split will vary with the system design. Systems in the EER range of 7.0 to 8.0 typically have design splits in the range 25 to 30° F. Systems in the EER range of 8.0 to 9.0 typically have design splits in the range 20 to 25° F. Systems with 9.0+EER ratings will have design splits in the range 15 to 20° F. The formula used for determining coil operating temperatures is: 
             COT   =       (       EAT   +   LAT     2     )     -   split           
where COT is the coil operating temperature, EAT is the entering air temperature of the coil (e.g., as measured by the temperature sensor  1026 ), LAT is the leaving air temperature of the coil (e.g., as measured by the temperature sensor  1022 ), and split is the design split temperature.
 
     The value (EAT+LAT)/2 is the average air temperature, which is also referred to as the mean temperature difference (MTD). It is also sometimes referred to as the coil TED or ΔT. 
     “Split” is the design split according to the EER rating. For example, a unit having an entering air condition of 80° F. DB and a 20° F. temperature drop across the evaporator  110  coil will have an operating coil temperature determined as follows: For an EER rating of 7.0 of 8.0: 
             COT   =         (       80   +   60     2     )     -     25   ⁢           ⁢   to   ⁢           ⁢     30   ∘         =     40   ⁢           ⁢   to   ⁢           ⁢     45     ∘               ⁢     F   .               
For an EER rating of 8.0 to 9.0:
 
             COT   =         (       80   +   60     2     )     -     20   ⁢           ⁢   to   ⁢           ⁢     25   ∘         =     45   ⁢           ⁢   to   ⁢           ⁢     50   ∘     ⁢           ⁢     F   .               
For an EER rating of 9.0+:
 
             COT   =         (       80   +   60     2     )     -     15   ⁢           ⁢   to   ⁢           ⁢     20   ∘         =     50   ⁢           ⁢   to   ⁢           ⁢     55   ∘     ⁢           ⁢     F   .               
Thus, the operating coil temperature changes with the EER rating of the unit.
 
     The surface area of the condenser  107  affects the condensing temperature the system  100  must develop to operate at rated capacity. The variation in the size of the condenser  107  also affects the production cost and price of the unit. The smaller the condenser  107 , the lower the efficiency (EER) rating. In the same EER ratings used for the evaporator  110 , at 95° outside ambient, the 7.0 to 8.0 EER category will operate in the 25 to 30° condenser  107  split range, the 8.0 to 9.0 EER category in the 20 to 25° condenser  107  split range, and the 9.0+EER category in the to 20° condenser  107  split range. 
     This means that when the air entering the condenser  107  is at 95° F., the formula for finding the condensing temperature is:
 
RCT=EAT+split
 
where RCT is the refrigerant condensing temperature, EAT is the entering air temperature of the condenser  107 , and split is the design temperature difference between the entering air temperature and the condensing temperatures of the hot high pressure vapor from the compressor  105 .
 
     For example, using the formula with 95° F. EAT, the split for the various EER systems would be: 
     For an EER rating of 7.0 to 8.0:
 
RCT=95°+25 to 30°=120 to 125° F.
 
     For an EER rating of 8.0 to 9.0:
 
RCT=95°+20 to 25°=115 to 120° F.
 
     For an EER rating of 9.0+:
 
RCT=95°+15 to 20°=110 to 115° F.
 
     The operating head pressures vary not only with changes in outdoor temperatures but also with the different EER ratings. 
     The amount of subcooling produced in the condenser  107  is determined primarily by the quantity of refrigerant in the system  100 . The temperature of the air entering the condenser  107  and the load in the evaporator  110  will have only a relatively small effect on the amount of subcooling produced. The amount of refrigerant in the system  100  has the predominant effect. Therefore, regardless of EER ratings, the unit should have, if properly charged, a liquid subcooled to 15 to 20° F. High ambient temperatures will produce the lower subcooled liquid because of the reduced quantity of refrigerant in the liquid state in the system. More refrigerant will stay in the vapor state to produce the higher pressure and condensing temperatures needed to eject the required amount of heat. 
     Table 1 shows 11 probable causes of trouble in an air conditioning system. After each probable cause is the reaction that the cause would have on the refrigeration system low side or suction pressure, the evaporator  110  superheat, the high side or discharge pressure, the amount of subcooling of the liquid leaving the condenser  107 , and the amperage draw of the condensing unit. In one embodiment, an airflow sensor (not shown) is included to measure the air over the condenser  107 . 
     Insufficient air over the evaporator  110  (as measured, for example, by using the airflow sensor  1023  and/or the differential pressure sensor  1025 ) is indicated by a greater than desired temperature drop in the air through the evaporator  110 . An unbalanced load on the evaporator  110  will also give the opposite indication, indicating that some of the circuits of the evaporator  110  are overloaded while others are lightly loaded. In one embodiment, the temperature sensor  1022  includes multiple sensors to measure the temperature across the evaporator  110 . The lightly loaded sections of the evaporator  110  allow liquid refrigerant to leave the coil and enter the suction manifold and suction line. 
     In TXV systems, the liquid refrigerant passing the sensing bulb of the TXV can cause the valve to close down. This reduces the operating temperature and capacity of the evaporator  110  as well as lowering the suction pressure. The evaporator  110  operating superheat can become very low because of the liquid leaving some of the sections of the evaporator  110 . 
     With inadequate airflow, high side or discharge pressure will be low due to the reduced load on the compressor  105 , reduced amount of refrigerant vapor pumped, and reduced heat load on the condenser  107 . Condenser  107  liquid subcooling would be on the high side of the normal range because of the reduction in refrigerant demand by the TXV. Condensing unit amperage draw would be down due to the reduced load. 
     In systems using fixed metering devices, the unbalanced load would produce a lower temperature drop of the air through the evaporator  110  because the amount of refrigerant supplied by the fixed metering device would not be reduced; therefore, the system pressure (boiling point) would be approximately the same. 
     The evaporator  110  superheat would drop to zero with liquid refrigerant flooding into the suction line  111 . Under extreme case of imbalance, liquid returning to the compressor  105  could cause damage to the compressor  105 . The reduction in heat gathered in the evaporator  110  and the lowering of the refrigerant vapor to the compressor  105  will lower the load on the compressor  105 . The compressor  105  discharge pressure (hot gas pressure) will be reduced. 
     The flow rate of the refrigerant will be only slightly reduced because of the lower head pressure. The subcooling of the refrigerant will be in the normal range. The amperage draw of the condensing unit will be slightly lower because of the reduced load on the compressor  105  and reduction in head pressure. 
     In the case of excessive load, the opposite effect exists. The temperature drop of the air through the coil will be less, because the unit cannot cool the air as much as it should. Air is moving through the coil at too high a velocity. There is also the possibility that the temperature of the air entering the coil is higher than the return air from the conditioned area. This could be from air leaks in the return duct system drawing hot air from unconditioned areas. 
     The excessive load raises the suction pressure. The refrigerant is evaporating at a rate faster than the pumping rate of the compressor  105 . If the system uses a TXV, the superheat will be normal to slightly high. The valve will operate at a higher flow rate to attempt to maintain superheat settings. If the system  100  uses fixed metering devices, the superheat will be high. The fixed metering devices cannot feed enough increase in refrigerant quantity to keep the evaporator  110  fully active. 
     The high side or discharge pressure will be high. The compressor  105  will pump more vapor because of the increase in suction pressure. The condenser  107  must handle more heat and will develop a higher condensing temperature to eject the additional heat. A higher condensing temperature means a greater high side pressure. The quantity of liquid in the system has not changed, nor is the refrigerant flow restricted. The liquid subcooling will be in the normal range. The amperage draw of the unit will be high because of the additional load on the compressor  105 . 
     When the temperature of the ambient air entering the condenser  107  is low, then the condenser  107  heat transfer rate is excessive, producing an excessively low discharge pressure. As a result, the suction pressure will be low because the amount of refrigerant through the metering device  109  will be reduced. This reduction will reduce the amount of liquid refrigerant supplied to the evaporator  110 . The coil will produce less vapor and the suction pressure drops. 
     The decrease in the refrigerant flow rate into the coil reduces the amount of active coil, and a higher superheat results. In addition, the reduced system capacity will decrease the amount of heat removed from the air. There will be higher temperature and relative humidity in the conditioned area and the high side pressure will be low. This starts a reduction in system capacity. The amount of subcooling of the liquid will be in the normal range. The quantity of liquid in the condenser  107  will be higher, but the heat transfer rate of the evaporator  110  is less. The amperage draw of the condensing unit will be less because the compressor  105  is doing less work. 
     The amount of drop in the condenser  107  ambient air temperature that the air conditioning system will tolerate depends on the type of pressure reducing device in the system. Systems using fixed metering devices will have a gradual reduction in capacity as the outside ambient drops from 95° F. This gradual reduction occurs down to 65° F. Below this temperature the capacity loss is drastic, and some means of maintaining head pressure must be employed to prevent the evaporator  110  temperature from dropping below freezing. Some systems control air through the condenser  107  via dampers in the airstream or a variable speed condenser  107  fan. 
     Systems that use TXV will maintain higher capacity down to an ambient temperature of 47° F. Below this temperature, controls must be used. The control of airflow through the condenser  107  using dampers or the condenser  107  fan speed control can also be used. In larger TXV systems, liquid quantity in the condenser  107  is used to control head pressure. 
     The higher the temperature of the air entering the condenser  107 , the higher the condensing temperature of the refrigerant vapor to eject the heat in the vapor. The higher the condensing temperature, the higher the head pressure. The suction pressure will be high for two reasons: (1) the pumping efficiency of the compressor  105  will be less; and (2) the higher temperature of the liquid will increase the amount of flash gas in the metering device  109 , further reducing the system efficiency. 
     The amount of superheat produced in the coil will be different in a TXV system and a fixed metering device system. In the TXV system the valve will maintain superheat close to the limits of its adjustment range even though the actual temperatures involved will be higher. In a fixed metering device system, the amount of superheat produced in the coil is the reverse of the temperature of the air through the condenser  107 . The flow rate through the fixed metering devices are directly affected by the head pressure. The higher the air temperature, the higher the head pressure and the higher the flow rate. As a result of the higher flow rate, the subcooling is lower. 
     Table 2 shows the superheat that will be developed in a properly charged air conditioning system using fixed metering devices. The head pressure will be high at the higher ambient temperatures because of the higher condensing temperatures required. The condenser  107  liquid subcooling will be in the lower portion of the normal range. The amount of liquid refrigerant in the condenser  107  will be reduced slightly because more will stay in the vapor state to produce the higher pressure and condensing temperature. The amperage draw of the condensing unit will be high. 
     
       
         
               
               
               
             
               
               
               
             
           
               
                   
                 TABLE 2 
               
               
                   
                   
               
               
                   
                 Air Temperature Entering 
                   
               
               
                   
                 Condenser 107 (° F.) 
                 Superheat (° F.) 
               
               
                   
                   
               
             
             
               
                   
               
             
          
           
               
                   
                 65 
                 30 
               
               
                   
                 75 
                 25 
               
               
                   
                 80 
                 20 
               
               
                   
                 85 
                 18 
               
               
                   
                 90 
                 15 
               
               
                   
                 95 
                 10 
               
               
                   
                 105 &amp; above 
                 5 
               
               
                   
                   
               
             
          
         
       
     
     A shortage of refrigerant in the system means less liquid refrigerant in the evaporator  110  to pick up heat, and lower suction pressure. The smaller quantity of liquid supplied to the evaporator  110  means less active surface in the coil for vaporizing the liquid refrigerant, and more surface to raise vapor temperature. The superheat will be high. There will be less vapor for the compressor  105  to handle and less heat for the condenser  107  to reject, lower high side pressure, and lower condensing temperature. The compressor  105  in an air conditioning system is cooled primarily by the cool returning suction gas. Compressors  105  that are low on charge can have a much higher operating temperature. 
     The amount of subcooling will be below normal to none, depending on the amount of undercharge. The system operation is usually not affected very seriously until the subcooling is zero and hot gas starts to leave the condenser  107 , together with the liquid refrigerant. The amperage draw of the condensing unit will be slightly less than normal. 
     An overcharge of refrigerant will affect the system  100  in different ways, depending on the pressure reducing device used in the system  100  and the amount of overcharge. 
     In systems using a TXV, the valve will attempt to control the refrigerant flow in the coil to maintain the superheat setting of the valve. However, the extra refrigerant will back up into the condenser  107 , occupying some of the heat transfer area that would otherwise be available for condensing. As a result, the discharge pressure will be slightly higher than normal, the liquid subcooling will be high, and the unit amperage draw will be high. The suction pressure and evaporator  110  superheat will be normal. Excessive overcharging will cause even higher head pressure, and hunting of the TXV. 
     For TXV systems with excessive overcharge the suction pressure will typically be high. Not only does the reduction in compressor  105  capacity (due to higher head pressure) raise the suction pressure, but the higher pressure will cause the TXV valve to overfeed on its opening stroke. This will cause a wider range of hunting of the valve. The evaporator  110  superheat will be very erratic from the low normal range to liquid out of the coil. The high side or discharge pressure will be extremely high. Subcooling of the liquid will also be high because of the excessive liquid in the condenser  107 . The condensing unit amperage draw will be higher because of the extreme load on the compressor  105  motor. 
     The amount of refrigerant in the fixed metering system has a direct effect on system performance. An overcharge has a greater effect than an undercharge, but both affect system performance, efficiency (EER), and operating cost. 
       FIGS. 12 through 14  show how the performance of a typical capillary tube air conditioning system is affected by an incorrect amount of refrigerant charge. In  FIG. 12 , at 100% of correct charge (55 oz), the unit develops a net capacity of 26,200 Btu/hr. When the amount of charge is varied 5% in either direction, the capacity drops as the charge is varied. Removing 5% (3 oz) of refrigerant reduces the net capacity to 25,000 Btu/hr. Another 5% (2.5 oz) reduces the capacity to 22,000 Btu/hr. From there on the reduction in capacity became very drastic: 85% (8 oz), 18,000 Btu/hr; 80% (11 oz), 13,000 Btu/hr; and 75% (14 oz), 8000 Btu/hr. 
     Overcharge has a similar effect but at a greater reduction rate. The addition of 3 oz of refrigerant (5%) reduces the net capacity to 24,600 Btu/hr; 6 oz added (10%) reduces the capacity to 19,000 Btu/hr; and 8 oz added (15%) drops the capacity to 11,000 Btu/hr. This shows that overcharging of a unit has a greater effect per ounce of refrigerant than does undercharging. 
       FIG. 13  is a chart showing the amount of electrical energy the unit demands because of pressure created by the amount of refrigerant in the system as the refrigerant charge is varied. At 100% of charge (55 oz) the unit uses 32 kW. As the charge is reduced, the wattage demand also drops, to 29.6 kW at 95% (3 oz), to 27.6 kW at 90% (6.5 oz), to 25.7 kW at 85% (8 oz), to 25 kW at 80% (11 oz), and to 22.4 kW at 75% (14 oz short of correct charge). When the unit is overcharged, the power consumed also increases. At 3 oz (5% overcharge) the power consumed is 34.2 kW, at 6 oz (10% overcharge) 39.5 kW, and at 8 oz (15% overcharge), 48 kW. 
       FIG. 14  shows the efficiency of the unit (EER rating) based on the Btu/hr capacity of the system  100  versus the power consumed by the condensing unit. At correct charge (55 oz) the efficiency (EER rating) of the unit is 8.49. As the refrigerant is reduced, the EER rating drops to 8.22 at 9% of charge, to 7.97 at 90%, to 7.03 at 85%, to 5.2 at 80%, and to 3.57 at 75% of full refrigerant charge. When refrigerant is added, at 5% (3 oz) the EER rating drops to 7.19. At 10% (6 oz) the EER is 4.8, and at 15% overcharge (8 oz) the EER is 2.29. 
     The effect of overcharge produces a high suction pressure because the refrigerant flow to the evaporator  110  increases. Suction superheat decreases because of the additional quantity to the evaporator  110 . At approximately 8 to 10% of overcharge, the suction superheat becomes zero and liquid refrigerant will leave the evaporator  110 . This causes flooding of the compressor  105  and greatly increases the chance of compressor  105  failure. The high side or discharge pressure is high because of the extra refrigerant in the condenser  107 . Liquid subcooling is also high for the same reason. The power draw increases due to the greater amount of vapor pumped as well as the higher compressor  105  discharge pressure. 
     Restrictions in the liquid line  108  reduce the amount of refrigerant to the metering device  109 . Both TXV valve systems and fixed metering device systems will then operate with reduced refrigerant flow rate to the evaporator  110 . The following observations can be made of liquid line  108  restrictions. First, the suction pressure will be low because of the reduced amount of refrigerant to the evaporator  110 . The suction superheat will be high because of the reduced active portion of the coil, allowing more coil surface for increasing the vapor temperature as well as reducing the refrigerant boiling point. The high side or discharge pressure will be low because of the reduced load on the compressor  105 . Liquid subcooling will be high. The liquid refrigerant will accumulate in the condenser  107 . It cannot flow out at the proper rate because of the restriction. As a result, the liquid will cool more than desired. Finally, the amperage draw of the condensing unit will be low. 
     Either a plugged fixed metering device or plugged feeder tube between the TXV valve distributor and the coil will cause part of the coil to be inactive. The system  100  will then be operating with an undersized coil, resulting in low suction pressure because the coil capacity has been reduced. The suction superheat will be high in the fixed metering device systems. The reduced amount of vapor produced in the coil and resultant reduction in suction pressure will reduce compressor  105  capacity, head pressure, and the flow rate of the remaining active capillary tubes. The high side or discharge pressure will be low. 
     Liquid subcooling will be high; the liquid refrigerant will accumulate in the condenser  107 . The unit amperage draw will be low. 
     In TXV systems, a plugged feeder tube reduces the capacity of the coil. The coil cannot provide enough vapor to satisfy the pumping capacity of the compressor  105  and the suction pressure balances out at a low pressure. The superheat, however, will be in the normal range because the valve will adjust to the lower operating conditions and maintain the setting superheat range. The high side or discharge pressure will be low because of the reduced load on the compressor  105  and the condenser  107 . 
     Low suction and discharge pressures indicate a refrigerant shortage. The liquid subcooling is normal to slightly above normal. This indicates a surplus of refrigerant in the condenser  107 . Most of the refrigerant is in the coil, where the evaporation rate is low due to the higher operating pressure in the coil. The amperage draw of the condensing unit would be low because of the light load on the compressor  105 . 
     If the hot gas line  106  is restricted, then the high side or compressor  105  discharge pressure will be high if measured at the compressor  105  outlet or low if measured at the condenser  107  outlet or liquid line  108 . In either case, the compressor  105  current draw will be high. The suction pressure is high due to reduced pumping capacity of the compressor  105 . The evaporator  110  superheat is high because the suction pressure is high. The high side pressure is high when measured at the compressor  105  discharge or low when measured at the liquid line  108 . Liquid subcooling is in the high end of the normal range. Even with all of this, the compressor  105  amperage draw is above normal. All symptoms point to an extreme restriction in the hot gas line  106 . This problem is easily found when the discharge pressure is measured at the compressor  105  discharge. 
     When the measuring point is the liquid line  108  at the condenser  107  outlet, the facts are easily misinterpreted. High suction pressure and low discharge pressure will usually be interpreted as an inefficient compressor  105 . The amperage draw of the compressor  105  must be measured. The high amperage draw indicates that the compressor  105  is operating against a high discharge pressure. A restriction apparently exists between the outlet of the compressor  105  and the pressure measuring point. 
     When the compressor  105  will not pump the required amount of refrigerant vapor (e.g., because it is undersized, or is not working at rated capacity), the suction pressure will balance out higher than normal. The evaporator  110  superheat will be high. The high side or discharge pressure will be extremely low. Liquid subcooling will be low because not much heat will be in the condenser  107 . The condensing temperature will therefore be close to the entering air temperature. The amperage draw of the condensing unit will be extremely low, indicating that the compressor  105  is doing very little work. 
     The following formulas can be used by the systems  900 ,  1000  to calculate various operating parameters of the system  100  using data from one or more of the sensors shown in  FIG. 10 . 
     Power is:
 
Watts=volts×amps×PF
 
where PF is the power factor.
 
     Heat is:
 
Btu= W×ΔT  
 
     Specific heat is:
 
Btu= W×c×ΔT  
 
     Sensible heat added or removed from a substance is:
 
 Q=W×SH×ΔT  
 
     Latent heat added or removed from a substance is:
 
 Q=W×LH  
 
     The refrigeration effect is: 
             W   =     200   NRE           
where W is weight of refrigerant circulated per minute (e.g., lb/min), 200 Btu/min is the equivalent of 1 ton of refrigeration, and NRE is the net refrigerating effect (Btu/lb of refrigerant).
 
     The coefficient of performance (COP) is: 
     
       
         
           
             COP 
             = 
             
               refrigerating_effect 
               
                 heat_of 
                 ⁢ 
                 _compression 
               
             
           
         
       
     
     System capacity is:
 
 Q   t =4.45 ×CFM×Δh  
 
where Q t  is the total (sensible and latent) cooling being done, CFM is the airflow across the evaporator  110 , and Δh is the change of enthalpy of the air across the coil.
 
     Condensing temperature is:
 
RCT=EAT+split
 
where RCT is the refrigerant condensing temperature, EAT is the temperature of the air entering the condenser  107 , and split is the design temperature difference between the entering air temperature and the condensing temperatures of the hot high-pressure vapor from the compressor  105 .
 
     Net cooling capacity is:
 
 HC=HT−HM  
 
where HT is the heat transfer (gross capacity), HM is the motor heat, and HC is the net cooling capacity.
 
     Airflow rate of a system can be expressed as:
 
 Q=Q   s (1.08 ×TD )
 
where Q is the flow rate in CFM, Q s  is the sensible-heat load in Btu/hr, and TD is the dry bulb temperature difference in ° F.
 
     In a fan, airflow (CFM) is approximately related to rotation (rpm) as follows: 
     
       
         
           
             
               
                 CFM 
                 2 
               
               
                 CFM 
                 1 
               
             
             = 
             
               
                 rpm 
                 2 
               
               
                 rpm 
                 1 
               
             
           
         
       
     
     In a fan, pressure is approximately related to rotation as follows: 
     
       
         
           
             
               
                 SP 
                 2 
               
               
                 SP 
                 1 
               
             
             = 
             
               
                 ( 
                 
                   
                     rpm 
                     2 
                   
                   
                     rpm 
                     1 
                   
                 
                 ) 
               
               2 
             
           
         
       
     
     In a fan, work is approximately related to rotation as follows: 
     
       
         
           
             
               
                 Bhp 
                 2 
               
               
                 Bhp 
                 1 
               
             
             = 
             
               
                 ( 
                 
                   
                     rpm 
                     2 
                   
                   
                     rpm 
                     1 
                   
                 
                 ) 
               
               3 
             
           
         
       
     
     In one embodiment, the tachometer  1033  is provided to measure the rotational velocity of the fan  123 . In one embodiment, the tachometer  1032  is provided to measure the rotational velocity of the fan  122 . In one embodiment, the system  1000  uses one or more of the above fan equations to calculate desired fan rotation rates. In one embodiment, the system  1000  controls the speed of the fan  123  and/or the fan  122  to increase system efficiency. 
     The quantity of air used for cooling based on the sensible cooling is approximately:
 
 CFM=H   s ( TD× 1.08)
 
     The sensible heat removed is:
 
 Q   s =1.08 ×CFM×DBT  difference
 
     The latent heat removed is:
 
 Q   L =0.68 ×CFM×gr  moisture difference
 
     The total heat removed is:
 
 Q   t   =Q   s   +Q   L  
 
or
 
 Q   t =4.5 ×CFM ×total heat difference
 
     The rate of heat transfer is:
 
 Q=U×A×TD  
 
where Q is the heat transfer (Btuh), U is the overall heat transfer coefficient (Btuh/ft 2 /° F.), A is the area (ft 2 ), TD is the temperature difference between inside and outside design temperature and the refrigerated space design temperature.
 
     The keypad  1050  is used to provide control inputs to the efficiency monitoring system. The display  1008  provides feedback to the user, temperature set point display. In one embodiment, the power use and/or power cost can be displayed on the display  1008 . In one embodiment, the system  1000  receives rate information from the power company to use in calculating power costs. In one embodiment, the absolute efficiency of the refrigerant-cycle system can be shown on the display  1008 . In one embodiment, the relative efficiency of the refrigerant-cycle system can be shown on the display  1008 . In one embodiment, the data from various sensors in the system  1000  can be shown on the display  1008 . In one embodiment, diagnostic messages (e.g., change the filter, add refrigerant, etc.) are shown on the display  1008 . In one embodiment, messages from the power company are shown on the display  1008 . In one embodiment, warning messages from the power company are shown on the display  1008 . In one embodiment, the thermostat  1001  communicates with the power company (or other remote device) using power line communication methods such as, for example, BPL. 
     When the system  1000  is configured, the installer programs in the fixed system parameters needed for calculation of efficiency and/or other quantities derived from the sensor data. Typical fixed programmed parameters include the type of refrigerant, the compressor specifications, the condenser specifications, the evaporator specifications, the duct specifications, the fan specifications, the system SEER, and/or other system parameters. Typical fixed programmed parameters can also include equipment model and/or serial numbers, manufacturer data, engineering data, etc. 
     In one embodiment, the system  1000  is configured by bringing the refrigerant-cycle system up to design specifications, and then running the system  1000  in a calibration mode wherein the system  1000  takes sensor readings to measure normal baseline parameters for the refrigerant-cycle system. Using the measured baseline data, the system  1000  can calculate various system parameters (e.g., split temperatures, etc.). 
     In one embodiment, the system  1000  is first run in a calibration mode to measure baseline data, and then run in a normal monitoring mode wherein it compares operation of the refrigerant-cycle system with the baseline data. The system  1000  then gives alerts to potential problems when the operating parameters vary too much from the baseline data. 
     In one embodiment, the system  1000  is configured by using a combination of programmed parameters (e.g., refrigerant type, temperature splits, etc.) and baseline data obtained by operating the refrigerant-cycle system. 
       FIG. 15  shows a differential-pressure sensor  1502  used to monitor an air filter  1501  in an air-handler system. As the air filter  1501  becomes clogged, the differential pressure across the air filter  1501  will rise. This increase in differential pressure is measured by the differential pressure sensor  1502 . The differential pressure measured by the differential pressure sensor  1502  is used to assess the state of the air filter  1501 . When the differential pressure is too high, then replacement of the air filter  1501  is indicated. 
       FIG. 16  shows the differential-pressure sensor  1502  from  FIG. 15  provided to a wireless communication unit to allow the data from the differential pressure sensor  1502  to be provided to other aspects of the monitoring system, such as, for example, the condenser unit sender  1002  or the thermostat  1001 . 
       FIG. 17  shows the system of  FIG. 16  implemented using a filter frame  1701  to facilitate retrofitting of existing air handler systems. The frame  1701  includes the sensor  1502  and a sender  1601 . The frame  1701  is configured to fit into a standard filter frame. The frame  1701  is configured to hold a standard filter  1501 . In one embodiment, the frame  1701  evaluates the cleanliness of the filter  1501  by measuring a differential pressure between the filter input and output air. In one embodiment, the frame  1701  evaluates the cleanliness of the filter  1501  by providing a source of light on one side of the filter, a light sensor on the other side of the filter, and by measuring the light transmission through the filter. In one embodiment, the frame  1701  is calibrated to a baseline light transmission level. In one embodiment, the frame  1701  signals that the filter is dirty when the light transmission falls below a fixed threshold level. In one embodiment, the frame  1701  calibrates a baseline light transmission level each time a clean, filter is installed. In one embodiment, the frame  1701  signals that the filter is dirty when the light transmission falls below a percentage of the baseline level. 
     Although various embodiments have been described above, other embodiments will be within the skill of one of ordinary skill in the art. Thus, for example, although described primarily in terms of an air-conditioning system, one of ordinary skill in the art will recognize that all or part of the system  1000  can be applied to other refrigerant-cycle systems, such as commercial HVAC systems, refrigerator systems, freezers, water chillers, etc. Thus, the invention is limited only by the claims that follow.