MX2007001671A

MX2007001671A - Method and apparatus for monitoring refrigerant-cycle systems.

Info

Publication number: MX2007001671A
Application number: MX2007001671A
Authority: MX
Inventors: Lawrence Kates
Original assignee: Lawrence Kates
Priority date: 2004-08-11
Filing date: 2005-06-27
Publication date: 2007-04-23
Also published as: EP1914481A3; EP1914482A3; EP1914482A2; EP1914483A2; EP1914481A2; EP1914483A3; RU2007108788A

Abstract

A real-time monitoring system that monitors various aspects of the operationof a refrigerant-cycle system is described. In one embodiment, the system includesa processor that measures power provided to the refrigerant-cycle system andthat gathers data from one or more sensors and uses the sensor data to calculatea figure of merit related to the efficiency of the system. In one embodiment, thesensors include one or more of the following sensors: a suction line temperaturesensor, a suction line pressure sensor, a suction line flow sensor, a hot gas linetemperature sensor, a hot gas line pressure sensor, a hot gas line flow sensor,a liquid line temperature sensor, a liquid line pressure sensor, a liquid lineflow sensor. In one embodiment, the sensors include one or more of an evaporatorair temperature input sensor, an evaporator air temperature output sensor,an evaporator airflow sensor, an evaporator air humidity sensor, and a differentialpressure sensor. In one embodiment, the sensors include one or more of a condenserair temperature input sensor, a condenser air temperature output sensor, anda condenser air flow sensor, an evaporator air humidity sensor. In one embodiment,the sensors include one or more of an ambient air sensor and an ambient humiditysensor.

Description

METHOD AND APPARATUS FOR VERIFYING REFRIGERANT CYCLE SYSTEMS FIELD OF THE INVENTION The present invention is concerned with a verification system for measuring the operation and efficiency of a refrigerant cycle system, such as, for example, an air conditioning system or system of refrigeration.

BACKGROUND OF THE INVENTION One of the major recurring expenses in the operation of a domestic 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 putting the system into operation is unnecessarily increased. Every pound of refrigerant circulating in the system must do its part of the work. It must absorb a quantity of heat in the evaporator or cooling coil and must dissipate this heat - plus something that is added to the compressor - by means of the condenser, either cooled by air, cooled by water or cooled by evaporation. The work performed by each pound of the refrigerant as it passes through the evaporator is reflected by the amount of heat it collects from the refrigeration load, mainly when the refrigerant undergoes a change in state from liquid to vapor.

For a liquid to be able to change to steam, heat must be added or absorbed in it. This is what happens in the cooling coil. The refrigerant enters the measuring device as a liquid and passes through the device to the evaporator, where it absorbs heat as it evaporates to vapor. As steam, it advances through the pipe or suction tube to the compressor. There it is compressed from a low temperature, low pressure steam to a high temperature, high pressure steam; then it passes through the high pressure tube or discharge tube to the condenser, where it undergoes another change of state - from vapor to liquid - in which state it flows out to the liquid tube and again advances to the measuring device for another trip through the evaporator. When the refrigerant as liquid exits the condenser it can advance to a receiver until it is needed in the evaporator or it can go directly to the liquid line to the measuring device and then to the evaporator coil. The liquid entering the measuring 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 the refrigerant tables in the Appendix. The steam leaving the evaporator will also have a latent 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 that is done per pound of refrigerant as it passes through the evaporator and absorbs heat. The amount of heat absorbed per pound of refrigerant is known as the cooling effect of the system or refrigerant within the system. Situations that can reduce the overall efficiency of the system include, refrigerant overload, refrigerant underload, refrigerant line restrictions, defective compressor, excessive load, insufficient load, undersized or dirty pipe, clogged air filters, etc. Unfortunately, modern HVAC systems do not include verification systems to verify the operation of the system. A modern HVAC system is commonly installed, charged with refrigerant by a service technician and then put into operation for months or years without additional maintenance. While the system is providing cold air, the owner of the building or owner of the house assumes that the system is working properly. This assumption can be expensive; since the owner has no knowledge of how well the system is working. If the efficiency of the system deteriorates, the system may still be able to produce the desired amount of cold air, but you will have to work harder and consume more energy to do this. In many cases, the system owner does not inspect or service the HVAC system until the efficiency has dropped so low that it can no longer cool the building. This is due in part because the service of an HVAC system requires specialized tools and knowledge that the owner of the typical building or owner of the house does not own. Thus, the owner of the building or owner of the house, must pay an expensive service call in order to evaluate the system. Even if the owner calls for a service call, many HVAC service technicians do not measure the system efficiently. Commonly, HVAC service technicians are trained only to perform rudimentary system checks (eg, refrigerant charge, exit temperature), but such rudimentary inspections may not discover other factors that can result in poor system efficiency. Thus, the owner of the building or owner of the typical house, puts into operation the HVAC system year after year not knowing that the system may be wasting money when put into operation at an efficiency lower than the peak. In addition, inefficiency in the use of electrical energy can lead to shutdowns and blackouts during heat waves or other periods of high air conditioning use due to the overload of the electric power system (commonly referred to as the national electric power system).

BRIEF DESCRIPTION OF THE INVENTION These and other problems are solved by a real-time verification system that verifies various aspects of the operation of a refrigerant system, such as for example an HVAC system, a refrigerator, a cooler, a freezer, a water cooler, etc. In one embodiment, the verification system is configured with a retroactive update system that can be installed in an existing refrigerant system. In one embodiment, the system includes a processor that measures the energy provided to the HVAC system and collects data from one or more detectors and uses the detector data to calculate a merit number related to the efficiency of the system. In one embodiment, the detectors include one or more of the following detectors: a suction line temperature detector, a suction line pressure detector, a suction line flow detector, a gas line temperature detector hot, a hot gas line pressure detector, a hot gas line flow detector, a liquid line temperature detector, a liquid line pressure detector, a liquid line flow detector. In one embodiment, the detectors include one or more of an evaporator air temperature input detector, an evaporator air temperature output detector, an evaporator air flow detector, an evaporator air humidity detector and a differential pressure detector. In one embodiment, the detectors include one or more of a condenser air temperature input detector, a condenser air temperature output detector and a condenser air flow detector, an evaporator air humidity detector . In one embodiment, the detectors include one or more of an ambient air detector and an ambient humidity detector.

BRIEF DESCRIPTION OF THE FIGURES Figure 1 is a diagram of a typical refrigerant cycle system used in HVAC systems, refrigerators, freezers and the like. Figure 2 is a detailed pressure-heat diagram of a typical refrigerant (R-22). Figure 3 is a pressure-heat diagram showing pressure-enthalpy changes through a refrigeration cycle. Figure 4 is a pressure-heat diagram showing pressure, heat and temperature values for a refrigeration cycle operating with an evaporator at 4.4 ° C (40 ° F). Figure 5 is a pressure-heat diagram showing pressure, heat and temperature values for a refrigeration cycle operating with an evaporator at -6.7 ° C (20 ° F).

Figure 6 is a pressure-heat diagram showing the cycle of Figure 4 with an evaporator temperature of 4.4 ° C (40 ° F), where the condensation temperature has increased to 48.9 ° C (120 ° F) ). Figure 7 is a pressure-heat diagram showing how the subcooling by the condenser improves the cooling effect and the COP. Figure 8 is a pressure-heat diagram showing the cooling process in the evaporator. Figure 9A is a block diagram of a verification system for verifying the operation of the refrigerant cycle system. 9B is a block diagram of a verification system for verifying the operation of the refrigerant cycle system, wherein operation data for the system is provided to a verification service, such as, for example, an electric power company or verification center, when using data transmission in power lines. Figure 9C is a block diagram of a verification system for verifying the operation of the refrigerant cycle system, where operation data for the system is provided to a verification service, such as for example an energy company or center of verification, when using data transmission in a computer network. Figure 9D is a block diagram of a verification system to verify the operation of the refrigerant cycle system, wherein data regarding the operation of the system is provided to a thermostat and / or a computer system, such as for example a site verification computer, a maintenance computer, a personal digital assistant, a personal computer, etc. . Figure 9E is a block diagram of a verification system for verifying the operation of the refrigerant cycle system, wherein an electronically controlled measuring device is provided to allow control of the system in an energy efficient manner. Figure 9F is a block diagram of a thermostat control and verification system having a data interface device provided to the thermostat. Figure 9G is a block diagram of a thermostat control and verification system having a data interface device provided to the evaporator unit. Figure 9H is a block diagram of a thermostat control and verification system having a data interface device provided to the condenser unit. Figure 10 (consisting of Figures 10A and 10B) shows several detectors that can be used in connection with the system of Figures 9A-H to verify the operation of the refrigerant cycle system. Figure 11 shows the temperature drop in the air through the evaporator as a function of humidity. Figure 12 shows the thermal capacity of a typical refrigerant cycle system as a function of the refrigerant charge. Figure 13 shows the energy consumed in a typical refrigerant cycle system as a function of the refrigerant charge. Figure 14 shows the efficiency of a typical refrigerant cycle system as a function of the refrigerant charge. Figure 15 shows a differential pressure detector used to verify an air filter in an air handling system. Figure 16 shows a differential pressure detector used to verify an air filter in an air handling system that uses a wireless system to provide filter differential pressure data back to other aspects of the verification system. Figure 17 shows the system of Figure 16 implemented using a filter frame to facilitate retroactive updating of existing air handling systems.

DETAILED DESCRIPTION OF THE INVENTION Figure 1 is a diagram of a typical 100 refrigerant cycle system used in HVAC systems, refrigerators, freezers and the like. In the system 100, a compressor provides hot compressed refrigerant gas to a hot gas line 106. The hot gas line supplies the hot gas to a condenser 107. The condenser 107 cools the gas and condenses and the gas to a liquid that is provided to a liquid line 108. The liquid refrigerant in the liquid line 108 is provided through a measuring device 109 to an evaporator 110. The refrigerant is again expanded to gas in the evaporator 110 and is provided back to the compressor through a suction line 110. A suction service valve 120 provides access to the suction line 111. A liquid line service valve 121 provides access to the liquid line 121. A fan 123 provides air of inlet 124 to the evaporator 110. The evaporator cools the air and provides the cooled evaporator outlet air 125. An optional dryer / accumulator 130 can be provided in the liquid line. 108. A fan 122 provides cooling air to the condenser 107. The measuring device 109 can be any refrigerant measuring device as used in the art, such as, for example, a capillary tube, a fixed orifice, a valve Thermostatic expansion (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 measuring device, such as a capillary tube or fixed orifice, will allow some adjustment in the capacity of the system as the load changes. As the condensing temperature on the outside increases, more refrigerant is fed through the measuring device to the evaporator, increasing its capacity slightly. Conversely, as the heat load goes down, the outdoor condensing temperature goes down and less refrigerant is fed to the evaporator. For a site where the load does not vary widely, fixed measuring devices can float with the load sufficiently well. However, for climates where there is a relatively greater range in temperature variation, an adjustable measuring device is commonly used. The systems 100 cools the air through the evaporator 110 by using the cooling effect of an expanding gas. This cooling effect is classified in Btu / lb of refrigerant (Btu / lb); If the total thermal load (given in Btu / h) is known, you can find the total number of pounds of refrigerant that must be circulated for each hour of system operation. This figure can be further decomposed to the amount that must be circulated every minute, by dividing the amount that is circulated per hour by 60. Due to a small hole in the measuring device 109, when the compressed refrigerant passes from the orifice more small in the measuring device to the largest tube in the evaporator, a change in pressure occurs along with a change in temperature. This temperature change occurs due to the vaporization of a small portion of the refrigerant (approximately 20%) and in the process of this vaporization, the heat that is involved is taken from the rest of the refrigerant. For example, from the saturated R-22 table in Figure 2, it can be seen that the liquid heat content at 37.8 ° C (100 ° F) is 39.27 BTU / lb and that liquid at 4.4 ° C (40 ° C). ° F) is 21.42 BTU / lb; this indicates that 17.85 BTU / lb have to be removed from every pound of refrigerant that enters the evaporator. The latent heat of vaporization of 4.4 ° C (40 ° F) (17.85 BTU / lb) is 68.87 BTU / lb. This is another method to calculate the cooling effect or work that is done, for each pound of refrigerant under the given conditions. The capacity of the compressor 105 must be such that it will remove from the evaporator that amount of refrigerant that has evaporated in the evaporator and in the measuring device in order to cause the necessary work to be carried out. The compressor 105 may be able to remove and send to the condenser 107 the same steam weight of the refrigerant, so that it can be condensed back to a liquid and thus continue in the cooling circuit 100 to perform additional work. If the compressor 105 is not capable of moving this weight, some of the steam will remain in the evaporator 110. This will in turn cause an increase in pressure inside the evaporator 110, accompanied by an increase in temperature and a decrease in work that It is done by the refrigerant and the design conditions within the refrigerated space can not be maintained. A compressor 105 that is too large will draw the refrigerant from the evaporator 110 too quickly, causing a decrease in temperature inside the evaporator 110, such that the design conditions will not be maintained. In order that the design conditions are maintained within a refrigeration circuit, a balance is maintained between the requirements of the evaporator 110 and the capacity of the compressor 105. This capacity is dependent on its displacement and its volumetric efficiency. The volumetric efficiency depends on the absolute suction and discharge pressures under which the compressor 105 is operating. In one embodiment, system 1000 controls the speed of compressor 105 to increase efficiency. In one embodiment, the system 1000 controls the measuring device 109 to increase efficiency. In one embodiment, system 1000 controls the speed of fan 123 to increase efficiency. In one embodiment, system 1000 controls the speed of fan 122 to increase efficiency. In system 100, the refrigerant passes from the liquid to the vapor state as it absorbs heat in the evaporator coil 110. In the stage of the compressor 105, the refrigerant vapor is increased in temperature and pressure, then the refrigerant gives up its heat in the condenser 107 to the room cooling medium and the refrigerant vapor condenses again to its liquid state where it is ready for use again in the cycle. Figure 2 shows the pressure, heat and temperature characteristics of this refrigerant. Enthalpy is another word for heat content. The diagrams such as figure 2 are referred to as pressure-enthalpy diagrams. Detailed pressure-enthalpy diagrams can be used to graph the cycle shown in figure 2, but a basic diagram or sketch as shown in figure 3 is useful as a preliminary illustration of the various phases of the refrigerant circuit. There are three basic areas in the graph denoting changes in state between the saturated liquid line 301 and the saturated steam line 302 in the center of the graph. The area to the left of the saturated liquid line 301 is the subcooled area, where the coolant has cooled below the boiling temperature corresponding to its pressure, while the area to the right of the saturated steam line 302 it is the overheating area, where the refrigerant vapor has been heated beyond the vaporization temperature corresponding to its pressure. The construction of diagram 300 illustrates what happens to the refrigerant in 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 placed on the graph, the other properties can be determined from the graph. If the point is located anywhere between the lines of saturated liquid 310 and steam 302, the refrigerant will be in the form of a mixture of liquid and vapor. If the site is closer to the saturated liquid line 301, the mixture will be more liquid than vapor and a point located from the center of the area at a particular pressure will indicate a situation of 50% liquid, 50% vapor. The change of state from vapor to liquid, the condensation process, occurs as the trajectory of the cycle develops from right to left; while the change of state from liquid to vapor, the process of evaporation, travels from left to right. The absolute pressure is indicated on the vertical axis on the left and the horizontal axis indicates the heat content or enthalpy, in BTU / lb. The distance between the two saturated lines 310, 302 at a given pressure, as indicated in the heat content line, adds the latent heat of vaporization of the refrigerant to the given absolute pressure. The distance between the two saturation lines is not the same at all pressures, since they do not follow parallel curves. Accordingly, there are variations in the latent heat of refrigerant vaporization, depending on the absolute pressure. There are also variations in the pressure-enthalpy graphs of different refrigerants and the variations depend on the various properties of the individual refrigerants. There is relatively little change in the temperature of the condensed cooling liquid after it leaves the condenser 107 and travels through the liquid line 108 in the direction of the expansion or measurement 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. Figure 4 shows the phases of the simple saturated cycle with appropriate marking of pressures, temperatures and heat content or enthalpy. Starting at point A in the saturated liquid, where all the refrigerant vapor at 37.8 ° C (100 ° F) has been condensed to liquid at 100 ° F and is at the entrance to the measuring device, between points A and B is the expansion process as the refrigerant passes through the measuring device 109 and the Coolant temperature is decreased from the condensing temperature of 37.8 ° C (100 ° F) to the evaporation temperature of 4.4 ° C (40 ° F). When vertical line A-B (the expansion process) is extended down to the lower axis, a reading of 39.27 BTU / lb is indicated which is the liquid heat content at 37.8 ° C (100 ° F). To the left of point B in the saturated liquid line 108 is the point Z, which is also at the temperature line of 4.4 ° C (40 ° F). Taking a vertical path down from point Z to the heat content line, a reading of 21.42 BTU / lb, which is the liquid heat content at 4.4 ° C (40 ° F), is indicated. The horizontal line between points B and C indicates the vaporization process in the evaporator 110, where the temperature at 4.4 ° C (40 ° F) absorbs enough heat to completely vaporize the refrigerant. Point C is to the saturated steam line, which indicates that the refrigerant has completely vaporized and is ready for the compression process. A line drawn vertically down to where it joins the enthalpy line indicates that the heat content, shown in hc is 108.14 Btu / lb, and the difference between ha and hc is 68.87 Btu / lb, which is the effect of refrigeration, as shown in the previous example. The difference between the points hz and hc in the enthalpy line totals 86.72 Btu / lb, which is the latent heat of vaporization of 1 Ib from R-22 to 4.4 ° C (40 ° F). This amount would also exhibit the cooling effect, but some of the refrigerant at 37.8 ° C (100 ° F) must be evaporated or vaporized so that the remaining portion of each pound of R-22 can be decreased at a temperature of 37.8 °. C (100 ° F) at 4.4 ° C (40 ° F). All refrigerants exhibit properties of volume, temperature, pressure, enthalpy or content of heat and entropy when in the gaseous state. Entropy is defined as the degree of disorder of the molecules that compose it. In refrigeration, entropy is the ratio of the thermal content of the gas to its absolute temperature in degrees Ran in. The graph of pressure-enthalpy graphical the line of constant entropy, which remains the same provided that the gas is compressed and does not add or take heat from the outside. 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 refrigeration cycle studies, to plot the compression line either along or parallel to a constant entropy line. In Figure 5, the CD line denotes the compression process, in which the pressure and temperature of the vapor are increased from those in the evaporator 110 to that in the condenser 107, with the assumption that there has been no heat collection in the suction line 111 between the evaporator 110 and the compressor 105. For a condensing temperature of 37.8 ° C (100 ° F), a pressure gauge would read approximately 13.8 kg / cm2 gauge (196 psi); but the graph is plotted in absolute pressure and the atmospheric pressure of 1.03 kg / cm2 (14.7 pounds / in2) is added to the pounds / in2 gauge, making it really 14.8 kg / cm2 (210.61 pounds / in2 absolute). Point D on the absolute pressure line is equivalent to the condensing temperature of 37.8 ° C (100 ° F); it is not in the saturated steam line, it is on the right in the superheated area, at a junction of the line of absolute 14.8 kg / cm2 (210.61 pounds / in2 absolute), the constant entropy line of 4.4 ° C (40 °) F) and the temperature line of approximately 53.3 ° C (128 ° F). A line drawn vertically down from point D intersects with the heat content line at 118.68 Btu / lb, which is h¿, the difference between hc and ha is 10.54 Btu / lb - the compression heat that has been added steamed. This amount of heat is the thermal energy equivalent of the work done during the refrigeration compression cycle. This is the theoretical discharge temperature, assuming that saturated steam enters the cycle; in actual operation, the discharge temperature may be 11.1 ° C - 19.4 ° C (20 ° to 35 ° F) higher than theoretically predicted. This can be verified 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 that are pushed or compressed together in a narrow manner, called commonly compression heat. The D-E line denotes the amount of overheating that must be removed from the steam before the condensation process can begin. A line drawn vertically down from point E to point e in the heat content line indicates the distance hd-he, or heat that adds up to 6.54 Btu / lb, since the vapor heat content at 37.8 ° C (100 ° F) is 112.11 Btu / lb. This overheating is usually removed in the hot gas discharge line or in the upper portion of the condenser 107. During this process the vapor temperature is lowered to the condensation temperature.

The lines EA 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 37.8 ° C (100 ° F) and an absolute pressure of 14.8 kg / cm2 absolute (210.61 pounds / in2 absolute, the same temperature and pressure prevail at point A, but the refrigerant is now in a liquid state) At any other point on the EA line, the refrigerant is in a vapor liquid combination phase; The point A is closer by, the greater the amount of refrigerant that has condensed to its liquid state At point A, each pound of refrigerant is ready to advance to the refrigerant cycle again as it is needed for the removal of refrigerant. heat of the load in the evaporator 110. Two factors that determine the coefficient of performance (COP) of a refrigerant are the cooling effect and the compression heat.The equation can be written as Effect d e refrigeration COP = (1) Compressive heat Substituting the values, from the pressure-enthalpy diagram of the simple saturated cycle previously presented, the equation would be: hc - ha 68.87 COP = = = 6.53 hd - hc 10.54 The COP is a speed or The measure of the theoretical efficiency of a refrigeration cycle 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 spent in the compression process, the higher the COP of the cooling system. The pressure-enthalpy diagrams of Figure 4 and 5 show a comparison of two simple saturated cycles that have different evaporation temperatures, to bring several differences in other aspects of the cycle. In order that a rough mathematical calculation comparison can be made, the cycles shown in Figures 4 and 5 will have the same condensing temperature, but the evaporation temperature will be lowered to -6.7 ° C (20 ° F). The values of A, B, C, D and E of Figure 4 as the cycle are compared with those in Figure 5 (with an evaporator 110 at 11.1 ° C (20 ° F)). The effect of cooling, heat of compression and heat dissipated in 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, classified in Btu / lb. For the evaporation temperature cycle at 11.1 ° C (20 ° F) shown in Figure 5: Net cooling effect (hc <- ha) = 67.11 Btu / lb Heat compression (hd- - hc <) = 67.11 Btu / lb When comparing the data with those of the cycle with the figure 4 of evaporation temperature of 4.4 ° C (40 ° F), it is shown that there is a decrease in the net cooling effect (NRE) of 2.6% and a increase in compression heat of 16.7%. There will be some increase in superheat, which must be removed either in the hot gas line 106 or the upper portion of the condenser 107. This is the result of a decrease in the suction temperature, the condensing temperature remains the same. From equation 1, it follows that the weight of refrigerant to be circulated per ton of cooling, in a cycle with an operating temperature of -6.7 ° C (20 ° F) and a condensing temperature of 37.8 ° C (100 °) F), is 2.98 Ib / minute / ton: 2QQ (Btu fmk ?.}. 2Q0Bñ? ffim ffi? Witiilb 2.981b / min Circulating more refrigerant commonly involves either a larger compressor 105 or the same size of the compressor 105 operating at higher RPMs. Figure 6 shows the original cycle with an operating temperature of 4.4 ° C (40 ° F) but the condensation temperature has increased to 48.9 ° C (120 ° F). Again taking the specific data of the heat content or enthalpy line, it is found that for the condensation temperature cycle at 48.9 ° C (120 ° F) that ha = 45.71, hc = 108.14, hd = 122.01, and he = 112.78. Thus, the net cooling effect (hc - ha <; ) = 62.43 Btu / lb, the compression heat (hd- - hc) = 13.87 Btu / lb, and the superheat of the condenser 107 (hd < - he-) = 9.23 Btu / lb. Compared to the cycle that has the condensing temperature of 37.8 ° C (100 ° F) (Figure 4), the cycle can also be calculated by allowing the temperature of the condensation process to increase to 48.9 ° C (120 ° F) ) (as shown in figure 7). Figure 7 shows a decrease in NRE of 9.4%, an increase in compression heat of 31.6% and an increase in superheat to be removed either in the discharge line or in the upper portion of capacitor 107 of 40.5%. With an operating temperature of 4.4 ° C (40 ° F) and a condensing temperature of 48.9 ° C (120 ° F), the weight of the refrigerant to be circulated will be 3.2 pound / min / ton. This indicates that approximately 10% more refrigerant should be circulated to do the same amount of work as when the concentration temperature was 37.8 ° C (100 ° F). Both of these examples show that for the best efficiency of a system, the suction temperature should be as high as is feasible and the condensation temperature should be as low as feasible. Of course, there are limitations as to the extremes under which the system 100 can operate satisfactorily and other means of increasing efficiency must then be considered. The economics of the equipment (cost + operating performance) finally determine the feasibility interval. Referring to figure 8, after the condensation process has been completed and all the refrigerant vapor at 48.9 ° C (120 ° F) is in liquid state, if the liquid can be subcooled to point A 'on the 37.8 ° line C (100 ° F) (a difference of 11.1 ° C (20 ° F)), the NRE (hc - ha) will be increased 6.44 Btu / lb. This increase in the amount of heat absorbed in the evaporator 110 without an increase in the compression heat will increase the COP of the cycle, since there is no increase in the energy introduced to the compressor 105. This subcooling can take place as long as the liquid is present. temporarily in storage in the condenser 107 or receiver or some of the liquid heat may be dissipated at room temperature as it passes through the liquid tube to the measuring device. 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 reach the compressor 105 in a saturated condition. Overheating is added to the steam after the evaporation process has been completed, in the evaporator 110 and / or in the suction line 111, also as in the compressor 105. If this overheating is only added in the evaporator 110, it is doing some useful cooling; if too much heat is removed from the load or product, in addition to the heat that was removed during the evaporation process. However, if the steam is superheated in the suction line 111 located outside the conditioned space, no useful cooling is carried out; still 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 through the compressor 105 and a pressure drop occurs through the measuring device 109. Thus, the compressor 105 and the measuring device maintain the pressure difference between the condenser 107 and the evaporator 110. Thus, a cooling system can be divided into the high side and low side portions. The high side contains high pressure steam and liquid refrigerant and is the part of the system that rejects heat. The low side contains the low pressure liquid vapor and the refrigerant and is the side that absorbs heat. Heat is always trying to reach a state of equilibrium by flowing from a hotter object to a colder object. Heat only flows in one direction, from hotter to colder. The temperature difference (TD) is what allows heat to flow from one object to another. The greater the difference in temperature, the faster the flow of heat. For the high side of a refrigeration unit to reject heat, its temperature must be above the ambient temperature or surrounding temperature. In order for the evaporator 110 to absorb heat, its temperature must be lower than the ambient temperature of the surroundings. Two factors that affect the amount of heat between two objects are the difference in temperature and the mass of the two objects. The greater the temperature difference between the refrigerant coil (for example, condenser 107 or evaporator 110) and the surrounding air, the faster the heat transfer will be. The larger the size of the refrigerant coil, the greater the mass of refrigerant, which also increases the rate of heat transfer. Technicians may already be designing coils that have higher temperature differences or larger areas to increase the heat transfer rate. To increase energy efficiency, the systems are designed with larger coils because it is more efficient to have a lower temperature and a larger area to transfer heat. It takes less energy to produce a smaller pressure / temperature difference within a refrigeration system. The manufacturers of the new high-efficiency air-conditioning systems use this principle. The same principle can be applied to the coils of the evaporator 110. The temperature differences between the inlet air of the evaporator 124 and the outlet air of the evaporator 125 are lower than those in the previous systems. Older air conditioning systems, with lower efficiency, can have evaporation coils operating at an output temperature of 1.7 ° C (35 ° F), while the newer, higher efficiency evaporator 110 can operate in the 7.2 ° C output range (45 ° F). Both evaporators 110 can collect the same amount of heat provided that the higher temperature, higher efficiency coil has a larger area and therefore more mass of refrigerant is exposed to the air stream to absorb heat. The higher evaporator coil temperature can produce less dehumidification. In humid climates, dehumidification can be an important part of the total air conditioning. The selection of the correct equipment is important to ensure the operation of the system and to obtain desired energy efficiencies. Previously, it was common practice at many sites for installers to select an evaporator 110 of a different tonnage than the capacity of condenser unit 101. While this practice in the past may provide higher efficiency, for most systems designed Most technically today appropriate correspondence is usually obtained by using the manufacturer's specifications in order to provide an appropriate operation. Misadjusted 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 an unbalanced system may not receive proper cooling of the return refrigerant vapor. As a result, the temperature of the compressor 105 will be higher and this can reduce the life of the compressor 105. As the vapor from the refrigerant leaves the discharge side of a compressor 105, it enters the condenser 107. As this vapor travels through of the condenser 107, the heat of the coolant dissipates to the surrounding air through the pipes and fins. As the 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 been converted to liquid. The liquid flows from the outlet of the condenser 107 through the liquid line 108 to the measuring device 109. The high-pressure liquid refrigerant, high temperature, passes through the measuring device 109 where its temperature and pressure change. As the pressure and temperature change, some of the liquid refrigerant boils forming an instant evaporation gas. As this mixture of refrigerant, liquid and vapor flows through the evaporator 110, the heat is absorbed and the remaining liquid refrigerant changes to steam. At the outlet of the evaporator 110, the steam flows back through the suction line 111 to the compressor 105. The compressor 105 extracts this low pressure, low temperature steam and converts it to high temperature, high pressure steam, where the cycle begins again. An ideally sized and functional system 105 is one in which the last refrigerant vapor fragment changes to liquid at the end of the condenser 107 where the last fragment of liquid refrigerant changes to vapor at the end of the evaporator 110. However, due to that it is impossible to have a system operating in this state ideally, the units are designed to have some additional cooling, called subcooling, of liquid refrigerant to ensure that no steam leaves the condenser 107. Even a small amount of steam leaving the condenser 107 can significantly reduce the efficiency of the system 100. On the evaporator 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. Return of the liquid refrigerant to the compressor 105 it can damage the compressor 105. Systems that must operate under a wide range of temperature conditions will have difficulty maintaining the desired level of subcooling and overheating. There are two components that can be used in these systems to improve the level of efficiency and safety in operation. They are the receiver and the accumulator. The receiver is placed in the line 108 and keeps a bit of extra coolant in such a way that the system has enough for high loads on hot days. The accumulator is placed in the suction line 111 and traps any liquid refrigerant that would flow back to the compressor 105 on cold days with light loads. A liquid receiver may be located at the end of the outlet of the condenser 107 to collect liquid refrigerant. The liquid receiver allows the liquid to flow to the receiver and any steam collected in the receiver flows back to the condenser 107 to be converted back to liquid. The line connecting the receiver to the condenser 107 is called the condensate line and must be large enough in diameter to allow the liquid to flow to the receiver and steam to flow back to the condenser 107. The condensate line must also have a slope towards the receiver to allow the liquid refrigerant to flow freely from the capacitor 107 to 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. The receivers must be dimensioned in such a way that all the refrigerant charge can be stored in the receiver. Some cooling 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 the liquid refrigerant to be collected at the bottom of the accumulator and to 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. The outlet of the accumulator draws steam through a U-shaped tube or chamber. There is usually a small hole in the bottom of the tube or U-shaped chamber that allows the liquid refrigerant and oil to be attracted to the suction line. Without this small hole, the refrigerant oil would be collected in the accumulator and would not return to the compressor 105. This small orifice allows some liquid refrigerant to enter the suction line. However, it is such a small amount of liquid refrigerant that it boils rapidly, so that there is little danger of liquid refrigerant flowing to compressor 105. Accumulators are often found in heat pumps. During the change cycle, the liquid refrigerant can flow back out of the outer coil. This liquid refrigerant could cause the compressor 105 to be damaged if it were not for the accumulator, which blocks its return. The pressure-heat day of Figure 8 shows the cooling process in the evaporator 110. Initially, the high pressure liquid is usually subcooled 4.4-5.5 ° C (8-10 ° F) or more. When the subcooled liquid from point A flows through the expansion device 109, its pressure drops to the pressure of the evaporator 110. Approximately 20% of the liquid boils to gas, cooling the remaining liquid-gas mixture. Its total heat (enthalpy) at point B is relatively unchanged from A. No external thermal energy has been exchanged. From points B to C, the rest of the liquid boils, absorbing the heat that flows from the load of the evaporator 110 (air, water, etc.). At point C, all the liquid has evaporated and the refrigerant is vapor at the saturation temperature corresponding to the pressure of the evaporator 110. Subcooling increases the efficiency of the cycle and can prevent the instantaneous evaporation of the gas due to the loss of pressure of components, tube friction or height increase. Many smaller cooling systems are designed so that the expansion device controls the flow of the refrigerant, such that the evaporator 110 will heat the steam beyond saturated conditions and ensure that no liquid droplet enters and possibly damages the compressor 105. It is assumed in the present for the purpose of simplicity that there is no pressure drop across the evaporator 110. Actually there are pressure drops which would slightly displace the evaporation and condensation processes of the constant pressure line shown. If the evaporator 110 does not have to superheat the refrigerant vapor, it can produce more cooling capacity. In smaller systems, the difference is relatively small and it is more important to protect the compressor 105. In 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 thermal load at the same temperature difference as a properly sized evaporator. The new equilibrium point will be reached at a lower pressure and suction temperature. The load will be reduced and the pressure and discharge temperature will also be reduced. An undersized evaporator and a reduced heat load have both similar effect on the refrigerant cycle because both are removing less heat from the refrigerant. As the ambient temperature increases, the load on the evaporator increases. When the load in the evaporator increases, the pressures increase. The operating points move up and to the right in the pressure-heat curve. As the load in the evaporator decreases, the load in the evaporator decreases and the pressure decreases. The operating points on the pressure-heat curve move downwards. Thus, knowledge of the ambient temperature is useful to determine if the system 100 is operating efficiently. Figure 9A is a block diagram of a verification system 900 for verifying the operation of the refrigerant cycle system. In Figure 9A, one or more detectors of the condenser unit 901 measure the operating characteristics of the fragments of the condenser unit 101, one or more detectors of the evaporator unit 902 measure the operating characteristics of the evaporator unit 102 and one or more environmental detectors 903 measure the environmental conditions. The detector data of the detectors of the condenser unit 901, detectors of the evaporator unit 902 and detectors of the condenser unit 903 are provided to a processing system 904. The processing system 904 uses the detector data to calculate the efficiency of the system, identify potential performance problems, calculate the use of energy, etc. In one embodiment, the 904 processing system calculates the use of energy and energy costs due to the inefficient operation. In one embodiment, the processing system 904 schedules maintenance of the filter according to the elapsed time and / or use of the filter. In one embodiment, the 904 processing system identifies potential performance problems (for example, low air flow, insufficient or unbalanced load, excessive load, low ambient temperature, high ambient temperature, refrigerant underload, refrigerant overload, liquid line restriction, suction line restriction, water restriction hot gas line, inefficient compressor, etc.). In one embodiment, the 904 processing system provides graphs or tables of usage and energy costs. In one embodiment, the 904 processing system of the verification system provides graphs or tables of additional energy costs due to the 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 the thermostat 952 are combined. Figure 9B is a block diagram of the system 900, wherein operation data of the refrigerant cycle system is provided to a remote verification service 950, such as, for example, an energy supply company or verification center. In one embodiment, the system 900 provides operating data concerning the operating efficiency of the refrigerant cycle system to the remote unit 950. In one embodiment, the remote verification service provides operating efficiency data to a utility company. electric power or government agency. The data can be transmitted on power lines as shown in Figure 9B and / or by using data transmission in a data network (e.g., Internet, a wireless network, a cable modem network, a telephone network, etc. .) as shown in Figure 9B and as also shown and discussed in relation to Figures 9F-H. Figure 9D is a block diagram of a verification system for verifying the operation of the refrigerant cycle system, wherein data concerning the operation of the system is provided to a thermostat 952 and / or to a computer system 953, such as for example, a site verification computer, a maintenance computer, a personal digital assistant, a personal computer, etc. Figure 9E is a block diagram of a verification system for verifying the operation of the refrigerant cycle system, wherein an electronically controlled measuring device 960 is provided to allow control of the system in an energy efficient manner. Figure 9F is a block diagram of a thermostat control and verification system having a data interface device 955 provided to the thermostat 952. The thermostat 952 commonly communicates with a controller of the evaporator unit 953 using control wiring of relatively low voltage. The control unit 953 commonly 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 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 monitor 950. Figure 9G is a block diagram of a thermostat control and verification system, wherein a data interface device 956 is provided to the controller 954. The data interface device 956 allows the remote monitor 950 to communicate with the capacitor unit. In one embodiment, the data interface device 956 allows the remote monitor to read data from the detector of the capacitor unit 101. In one embodiment, the data interface device 956 allows the remote monitor to turn off the capacitor unit 101. In a modality, the data interface device 956 allows the remote monitor to change the compressor 105 to a lower speed mode. In one embodiment, the data interface device 956 allows the remote monitor to change or switch the capacitor unit 101 to an energy conservation mode. Figure 9H is a block diagram of a thermostat control and verification system where a data interface device 957 is provided to the controller 953. In one embodiment, the data interface devices 955-957 are configured as modems of electric line (for example, using broadband technology in electric line (BPL), or other network technology of power lines). 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 detector data from the detectors of the condenser unit 901 and / or the detectors of the evaporator unit 902 to the remote verification service 950. In one embodiment, the system 900 uses data from the detectors. of the condenser unit 901 and / or the detectors of the evaporator unit 902 to calculate an efficiency factor for the refrigerant cycle system and the system 900 provides the efficiency factor to the remote verification service 950. In one embodiment, System 900 provides energy usage data (eg, amount of energy used) by the refrigerant cycle system and system 900 provides the efficiency factor to remote verification service 950. In one embodiment, system 900 provides an identification code (ID) with the data transmitted to the 950 remote monitor to identify the 900 system. In one embodiment, the 950 remote monitor is provided with data regarding a maximum expected efficiency for the refrigerant cycle system (based on the manufacturing characteristics and design of the system). refrigerant cycle), in such a way that the remote monitor 950 can investigate the relative efficiency (that is, how the refrigerant cycle system is operating with respect to its expected operating efficiency). In one mode, the 950 remote monitor provides efficiency data to the utility or to a government agency in such a way that electric rates can be charged according to the efficiency of the system. In one embodiment, the owner of the house (or building owner) is charged with a higher electric rate for the electric power provided to a refrigerant cycle system that is operating at a relatively low absolute efficiency. In one mode, the home owner (or building owner) is charged a higher electric rate for the electric power provided to a refrigerant cycle system that is operating at a relatively low efficiency. In one embodiment, an electric rate is charged to the owner of the house (or building owner) according to a combination of relative and absolute efficiency of the refrigerant cycle system. In one embodiment, the data provided to the 950 verification system is used to provide notification 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 950 verification system is used to provide notification 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 mode, 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 950 system can remotely turn off the refrigerant cycle system by sending commands or commands to one or more of the interface devices 955-957. In one mode, the owner of the house is charged (or building owner) a higher electric rate for electric power provided to a refrigerant cycle system that is operating at a relatively low efficiency for a specific period of time, such as, for example, when the power system is highly charged, during periods of cooling after the peak half-day, during heat waves, during rolling blackouts, etc. In one embodiment, the home owner (or building owner) is charged a higher electric rate (a premium rate) for electric power provided to a refrigerant cycle system for a specific period of time, such as when The energy system is highly charged, during periods of cooling after half a peak day, during heat waves, during rolling blackouts, etc. In one mode, the homeowner (or building owner) can program the 900 system to receive messages from the electric utility that indicates premium rates are charged. In one mode, the homeowner (or building owner) can program the 900 system to shut down during periods of premium rates. In one mode, the homeowner (or building owner) can avoid paying premium rates by allowing the utility to remotely control the operation of the refrigerant cycle system during times of premium rates. In a modality, only the owner of the home (or building owner) is allowed to operate the refrigerant cycle system during periods of premium rates if the system is operating above a prescribed efficiency. In one embodiment, the system 900 verifies the amount of time that the refrigerant cycle system has been in operation (e.g., the amount of operating time during the last day, week, etc.). In one embodiment, the remote verification system may interrogate the system 900 to obtain data regarding the operation of the refrigerant cycle system and one or more of the data interface devices 955-957 will receive the interrogation and send the received data to the 950 verification system. The interrogation data are, for example, the efficiency classification of the refrigerant cycle system (for example, the SEER, EER, etc.), the current operating efficiency of the refrigerant cycle system, the operating time of the system for a specified period of time, etc. The 950 system operator (for example, the power company or power transmission company) can use the interrogation data to make load balancing decisions. Thus, for example, the decision as to whether to instruct the refrigerant cycle system to shut down or go to a low energy mode may be based on the efficiency of the system (specified efficiency, absolute efficiency and / or relative efficiency), the amount of time the system has been in operation, the willingness of the owner of the house or building to pay premium rates during periods of load-cutting, etc. So, for example, a homeowner who has a low efficiency system that is highly used or who has indicated a desire not to pay premium rates, would have their refrigerant cycle system turned off by the 950 system before that of a domestic owner who has installed a high efficiency system, which is used relatively little and which indicated the desire to pay premium rates. In one modality, when making the decision to shut down the 900 system, the 950 verification system would take into account the efficiency of the 900 system, the amount that the 900 system is used and the willingness of the owner 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 shut off during a power emergency) and slightly used systems are preferred. with respect to highly used systems). In one embodiment, system 900 sends data with respect to the set temperature of thermostat 052 to verification system 950. In one embodiment, the electricity rate charged to the owner of the house (or owner of the building) calculated according to a point adjustment of the 952 thermostat, such that a lower set point results in a higher rate load per kilowatt-hour. In one mode, the electricity rate charged to the home owner (or building owner) calculated in accordance with the thermostat setpoint 952 and the relative efficiency of the refrigerant cycle system in such a way that one more set point lower and / or lower efficiency results in a higher tariff load per kilowatt-hour. In one mode, the electricity rate charged to the home owner (or building owner) calculated in accordance with the thermostat setpoint 952 and the absolute efficiency of the refrigerant cycle system such that one set point more low and / or lower efficiency results in a higher tariff load per kilowatt-hour. In one embodiment, the electricity rate charged to the home owner (or building owner) calculated in accordance with the thermostat setpoint 952, the relative efficiency of the refrigerant cycle and the absolute efficiency of the refrigerant cycle system of 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 950 verification system can send instructions to the 900 system to shut down if the refrigerant cycle system is operating at low efficiency. In one embodiment, the verification system 950 can send instructions to the 9090 system to change the setting of the thermostat 952 (e.g., raise the set temperature of the thermostat 952) in response to the low efficiency of the refrigerant cycle system and / or to avoid a blackout. In one embodiment, the verification system can send instructions to the condenser unit 101 to switch the compressor 105 to a low speed mode to save energy. In one embodiment, the remote verification service knows the identification codes or addresses of the data interface devices 955-957 and correlates the identification codes with a database to determine if the refrigerant cycle system is servicing a relatively high priority client such as a hospital, the home of elderly or disabled people, etc. In such circumstances, the remote verification system can provide relatively less cooling cut off provided by the refrigerant cycle system. In a modality, system 900 communicates with verification system 950 to provide load clipping. Thus, for example, the verification system (e.g., an electric utility) can communicate with the data interface device 956 and / or the data interface device 957 to shut down the refrigerant cycle system. Thus, the 950 verification system can turn on and off times of air conditioners in a region to reduce the electrical load without implementing rolling blackouts. In one embodiment, the data interface device 956 is configured as a retroactive update device that can be installed in a capacitor unit to provide remote shutdown. In one embodiment, the data interface device 956 is configured as a retroactive update device that can be installed in a capacitor unit to remotely switch the capacitor unit to a low energy mode (e.g., energy saving) . In one embodiment, the data interface device 957 is configured as a retroactive update device that can be installed in an evaporator unit to provide remote shutdown or to remotely switch the system to a lower energy mode. In one embodiment, remote system 950 sends separate shutdown and restart commands to one or more of the data interface devices 955-957. In one embodiment, the remote system 950 sends commands to the data interface devices 955-957 to turn off for a specific period of time (eg, 10 minutes, 30 minutes, 1 hour, etc.) after which the system it restarts automatically. In one embodiment, the system 900 communicates with the verification system 950 to control the setting point of temperature setting of the thermostat 952 to prevent blackouts or partial blackouts regardless of the efficiency of the refrigerant cycle system. When there are conditions of partial blackouts or potential blackouts, the system 900 can cancel the thermostat setting of the homeowner to cause the temperature set point in the thermostat 952 to change (for example, increase (in order to reduce the use of energy In most residential installations, low-voltage control wiring is provided between the 952 thermostat and the evaporator unit 102 and the condenser unit 101. In most residential (and many industrial) applications ) the thermostat 952 receives electrical power via the low voltage control wiring of a down scaling transformer provided with the evaporator unit 102. In one embodiment, the modem 955 is provided in connection with the power meter 949 and the modem 955 communicates with the 952 thermostat using wireless communications In a typical air conditioning or cooling system ico, the condenser unit 101 is placed outside the area that is cooled and the evaporator unit 102 is calculated to the interior of the area that is cooled. The nature of exterior and interior depend on the particular installation. For example, in an air conditioning system or HVAC, the condenser unit 101 is commonly placed outside the building and the evaporator unit 102 is commonly 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 must be turned out (for example far away) from the area that is 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, System 900 is also programmed by specifying the size of the air handling system. In one embodiment, system 900 is also programmed by specifying the expected efficiency (eg design) of system 100. The verification system can do a better job of verifying efficiency than published performance classifications such as energy efficiency ratio (EER) and SEER. The EER is determined by dividing the steady state capacity published by the steady-state energy input published at 26.7 ° C (80 ° F) dry bulb / 19.4 ° C (67 ° F) indoors and 35 ° C (95 ° F) dry outdoor bulb.

This is a still not real objective with respect to the "real world" operating conditions of the system. The published SEER classification of a system is determined by multiplying the measured steady state EER at outdoor temperature conditions of 27.8 ° C (82 ° F), 26.7 ° C (80 ° F) dry bulb / 19.4 ° C (67 ° C) F) wet bulb of the inlet air temperature of the interior by the partial load factor (run time) (PLF) of the system. A major factor not considered in the SEER calculations is the actual partial load factor of the indoor evaporator cooling coil, which reduces the listed BTUH capacity of the unit and SEER efficiency level. Many older air handlers and duct systems do not provide published BTUH and SEER ratings. This is mainly due to the improper air flow through the evaporator 110, a dirty evaporator 110 and / or dirty fan wheels. Also, the improper location of supply diffusers and return air registers may result in inefficient floor level recirculation of the cold air conditioner, resulting in lack of heating load of the evaporator 110. When checking the system under conditions of actual load and when measuring the relevant ambient temperature and humidity, the system 900 can calculate the actual efficiency of the system 100 in operation. Figure 10 shows a verification system 1000 for verifying the operation of the refrigerant cycle system 100. The system 1000 shown in Figure 10 is an example of a system embodiment 900 shown in Figures 9A-E. In the system 1000, an emitter 1002 of the condenser unit verifies the operation of the condenser unit 101 by means of one or more detectors, an emitting unit 1003 of the evaporator verifies the operation of the evaporator unit 102 by means of one or more detectors. The emitter 1002 of the condenser unit and the emitter 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 Figure 10, processor 904 and thermostat 952 of Figures 9A-E are shown as a single thermostat-processor. Those of ordinary skill in the art will recognize that the processor functions can be separated from the thermostat. In one embodiment, a temperature detector is provided from the interior of building 1009 to thermostat 101. In one embodiment, an indoor humidity detector of building 1010 is provided to thermostat 101. In one embodiment, thermostat 1001 includes a display 1008 to display the status and efficiency of the system. In one embodiment, the thermostat 1001 includes a 1050 keypad and / or indicator lights (for example LEDs) 1051. A power detector 1011 for detecting the electrical energy consumed by the compressor 105 is provided to the emitter of the condenser unit 1002. In one embodiment, an energy detector 1017 for detecting the electrical energy consumed by the The condenser fan 122 is provided to the emitter of the condenser unit 1002. The air 125 of the evaporator 110 flows in the line 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, the emitter of the condenser unit 1002 is provided. 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 emitter of the condenser unit 1002. In one embodiment, a temperature detector 1014, configured to measure the temperature of the refrigerant in the fluid line 108 near the condenser 107, is provided to the emitter of the condenser unit 1002. The contaminants in the refrigerant lines 111, 106, 108, etc. They can reduce the efficiency of the refrigerant cycle system and can reduce the life of the compressor or other system components. In a modality, one or more pollutant detectors 1034, configured to detect contaminants in the refrigerant (e.g. water, oxygen, nitrogen, air, inappropriate oil, etc.) are provided in at least one of the refrigerant lines and provided to the emitter of the condenser unit 1002 (or optionally the emitter of the evaporator unit 1003). In one embodiment, the contaminant detector 1060 detects refrigerant fluid or droplets at the inlet to the compressor 105, which may cause damage to the compressor 105. In one embodiment, a contaminant detector 1060 is provided in the liquid line 108 for detecting bubbles in the refrigerant. The bubbles in the liquid line 106 may indicate low levels of refrigerant, an undersized capacitor 109, insufficient cooling of the capacitor 109, etc. In one embodiment, the detector 1034 detects water or steam in the refrigerant lines. In one embodiment, the detector 1034 detects acid in the refrigerant lines. In one embodiment, the detector 1034 detects acid in the refrigerant lines. In one embodiment, detector 1034 detects air or other gases (e.g., oxygen, nitrogen, carbon dioxide, chlorine, etc.). In one embodiment, a pressure sensor 1013, configured to measure the pressure in the suction line 111, is provided to the emitter of the condenser unit 1002. In one embodiment, a pressure sensor 1015, configured to measure the pressure in the liquid line 108, is provided to the emitter of the condenser unit 1002. In one embodiment, a pressure sensor (not shown), configured to measure the pressure in the hot gas line 106, is provided to the emitter of the unit. capacitor 1002. In one embodiment, pressure detector 1013 and pressure sensor 1015 are connected to system 100, by attaching pressure sensors 1013 and 1015 to service valves 120 and 121, respectively. Attaching the pressure sensors to the pressure valves is a convenient way to gain access to the refrigerant pressure in a retroactive update facility without having to open the pressurized refrigerant system. In one embodiment, a flow detector 1031, configured to measure flow in the suction line 111, is provided to the emitter of the condenser unit 1002. In one embodiment, a flow detector 1030, configured to measure the flow in the liquid line 108, is provided to the emitter of the condenser unit 1002. In one embodiment, a flow detector (not shown), configured to measure the flow in the hot gas line 106, is provided to the emitter of the unit. capacitor 1002. In one embodiment, the flow detectors are ultrasonic detectors 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 emitter of the condenser unit 1002. In one embodiment, a humidity detector 1029 configured to measure ambient humidity is provided to the emitter of the condenser unit. 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 emitter unit 1003. In one embodiment, a temperature detector 1021, configured to measuring the temperature of the refrigerant in the suction line 111 near the evaporator 110 is provided to the emitter unit 1003. In one embodiment, a temperature sensor 1026, configured to measure the temperature of the air 124 flowing to the evaporator 110 is provided to the emitter unit 1003. In one embodiment, a temperature detector 1026, configured to measure the temperature of the air 125 flowing out of the evaporator 110 is provided to the emitter unit 1003. In one embodiment, a flow detector 1023, configured to measure the air flow of the air 125 flowing out of the evaporator 110 is provided to the emitter unit 1003. In an embodiment, a moisture detector 1024, configured to measure the temperature of the air 125 flowing out of the evaporator 110 is provided to the emitter unit 1003. In one embodiment, a differential pressure detector 1025, configured to measure the pressure drop to through the evaporator 110, it is provided to the emitter unit 1003. In one embodiment, the temperature detectors are attached to the refrigerant lines (e.g., lines 106, 108, 111, in order to measure the temperature of the refrigerant In one embodiment, the temperature detectors 1012 and / or 1016 are provided to the interior of the compressor 105. In one embodiment, the temperature detectors are provided inside one or more of the refrigerant lines. A tachometer 1033 detects the rotational speed of the blades of the fan 123. The tachometer is provided to the emitter of the evaporator unit 1003. A tachometer 1032 detects the rotational speed of the fan blades on the condenser fan 122. The tachometer 1032 is provided to the emitter of the condenser unit 1002. In one embodiment, an energy detector 1027, configured to measure the electrical energy consumed by the fan 123 is provided to the unit. of transmitter 1003. In one embodiment, the emitter unit 1003 communicates data from the detector to the emitter of the capacitor unit 1002 via wireless transmission. In one embodiment, the emitter unit 1003 communicates data from the detector to the emitter of the condenser unit 1002 by means of existing HVAC wiring. In one embodiment, the emitter unit 1003 communicates detector data to the emitter of the capacitor unit 1002 by means of existing HVAC wiring by modulating the detector data on a carrier that is transmitted using the existing HVAC wiring. Each of the detectors shown in Figure 10 (for example, detectors 1010-1034 etc.) are optional. The system 1000 can be configured with a subset of the detectors illustrated in order to reduce the cost at the expense of the capacity of the verification system. Thus, for example, the contaminant detectors 1034 may be eliminated, but the ability of the system 1000 to detect contaminants detected by the detector 1034 will be compromised or lost. The pressure sensors 1013 and 1015 measure the suction and discharge pressure, respectively, in the compressor 105. The temperature sensors 1026 and 1022 measure the supply air of the evaporator 110 and the return air, respectively. The temperature sensors 1018 and 1019 measure the inlet air and discharge air, respectively, in the capacitor 107. The energy detectors 1011, 1017, and 1027 are configured to measure electrical energy. In one embodiment, one or more of the energy detectors measure the voltage supplied to a load and the energy is calculated by using a specified impedance for the load. In one embodiment, one or more of the energy detectors measure the current supplied to a load and the energy is calculated by using a specified impedance for the load. In one embodiment, one or more of the energy detectors measure voltage and current supplied to a load and the energy is calculated by using a specified power factor for the load. In one embodiment, the energy detectors measure voltage, current, and the phase relationship between voltage and current. The temperature sensors 1012 and / or 1021 measure the temperature of the refrigerant in the suction line 111. By measuring the temperature of the suction line 111, the superheat can be determined. The suction pressure has been measured by the pressure sensor 1013, the evaporation temperature can be read from a pressure-temperature graph. Overheating is the difference between the temperature of the suction line 111 and the evaporation temperature. The temperature sensors 1014 and / or 1020 measure the temperature of the refrigerant in the liquid line 108. When measuring the temperature of the liquid line 108, sub-cooling can be determined. The discharge pressure is measured by the pressure sensor 1015 and thus the condensation temperature can be read from the pressure-temperature graph. The subcooling is the difference between the temperature of the liquid line 108 and the condensation temperature. In one embodiment, the system 1000 calculates the efficiency when measuring the work (cooling) performed by the refrigerant cycle system and when dividing by the power consumed by the system. In one embodiment, system 1000 checks the system for abnormal operation. Thus, for example, in one embodiment, the system 1000 measures the temperature drop of the refrigerant through the condenser 109 using the temperature detectors 1016 and 1014 to be used to calculate the heat removed by the condenser. The system 1000 measures the temperature drop of the refrigerant through the evaporator 110 to be used to calculate the heat absorbed by the evaporator 110. The verification system is commonly used to verify the operation of a system 100 that was originally inspected and put into service. proper operation condition. The mechanical problems in an air conditioning system are generally classified into two categories: air side problems and cooling side problems.

The main problem that can occur in the air category is a reduction in air flow. Air handling systems do not suddenly increase in capacity, that is, they increase the amount of air through the coil. On the other hand, the cooling system does not suddenly increase in heat transfer capacity. 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 temperatures of the return air and supply air and subtract to obtain the temperature drop, the system 1000 Check to see the temperature difference is higher or lower than it should be. Figure 11 the temperature drop in the air through the evaporator as a function of humidity. In one embodiment, moisture detectors 1024 and / or 1041 are used to measure building humidity and / or humidity detector 1041 is used to measure ambient humidity. The humidity readings are used to correct the temperature readings for the wet bulb temperature according to the relative humidity. In one embodiment, a comparison of the desired (or expected) temperature drop across the evaporator 110 with the actual measured temperature drop is used to help classify potential air problems from the refrigerant cycle problems. If the actual temperature drop is less than the required temperature drop, then the air flow has probably been reduced. The reduced air flow can be caused by problems of dirty air filter or evaporator 110, with fan 123 and / or unusual restrictions in the duct system. Air filters of the pass type are commonly replaced at least twice a year, at the beginning of both the cooling and heating seasons. In one mode, the thermostat allows the owner to indicate when a new air filter is installed. The thermostat keeps track of the time the filter has been in use and provides a reminder to the owner when the filter needs to be replaced. In one embodiment, the thermostat uses a real elapsed clock time to determine the use of the filter. In one embodiment, the thermostat 1001 calculates the use of the filter according to the amount of time that the air handler has been blowing air through the filter. Thus, for example, in moderate climates or seasons or seasons where the air handling system is not used continuously, the thermostat will wait for a longer period of real time before indicating that the filter replacement is guaranteed. In some areas of higher use or where the dust content is high, the filter will generally have to be replaced relatively more frequently. In one embodiment, the thermostat uses a weighting factor to combine the operating time with the inactive time to determine the use of the filter. Thus, for example, to determine the use of the filter, hours when the air handler is blowing air through the filter are compensated relatively stronger than the hours at which the air handling system is inactive. In one embodiment, the owner can program the thermostat to indicate that the filter replacement is necessary after a specified number of hours or days (for example, actual days, such as days of operation or combination thereof). In one embodiment, the thermostat 1001 is configured to receive information from a source of information regarding daily atmospheric dust conditions and to use such information to calculate the use of the filter. Thus, in one embodiment, when calculating the use of the filter, the thermostat ponders days of relatively high atmospheric dust more strongly 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 internet, a pager network, a local area network, etc. In one embodiment, the thermostat collects data to calculate the filter usage and passes such data to a computer verification system. In commercial and industrial applications, a regular maintenance schedule is generally used. In one embodiment, detectors are provided in relation to the air filter, as described below in connection with Figure 11. In one embodiment, the energy measured by the energy meter 1027 is used to help diagnose and detect problems with the fan 123 and / or the air handling system. If the fan 123 is drawing too much or very little current or if the fan 123 is showing a low power factor, then possible problems with the fan and / or air handling system are indicated. Placing furniture or carpets over the return air grilles reduces the air available for the fan to handle. Air disruption to unused areas reduce air over the evaporator 110. Covering a return air grill to reduce the noise of the centrally located oven or air handler can reduce objectionable noise, but it also drastically affects the operation of the system by reducing the amount of air. The collapse of the return air duct system will affect the entire performance of the duct system. Air leaks in the return duct will raise the return air temperature and reduce the temperature drop across the coil.

The air flow detector 1023 can be used to measure the air flow through the conduits. In one embodiment, the air flow detector 1023 is a hot wire (or hot film) mass flow detector. In one embodiment, the differential pressure detector 1025 is used to measure the flow of air through the evaporator 110. In one embodiment, the differential pressure detector 1025 is used to measure the fall through the evaporator 110. In one embodiment, the pressure drop across the evaporator is used to estimate when the evaporator 110 is restricting the air flow (for example, due to damage, dirt, hair, dust, etc.). In one embodiment, the differential pressure detector 1025 is used to measure the fall through an air filter to estimate when the filter is restricting air flow (eg, 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. 1023 is used to measure the air flow to the pipes 1080. In one embodiment, the indicator lights 1051 are used to indicate that the air flow to the 1080 pipeline is restricted (eg, due to dirt, furniture or carpets placed in front of vents, closed vents, dirty evaporator, dirty fan blades, etc.) In one embodiment, a dust detector is provided in the air stream of the evaporator 110. In one embodiment, the dust detector includes a light source (optical and / or infrared) and a light detector. The dust detector measures the transmission of light between the source and the light detector. The accumulation of dust will cause the light to be attenuated. The detector detects the presence of dust accumulated in the evaporator 110 by measuring the attenuation of light between the light source and the light detector. When the attenuation exceeds a desired value, the verification system 1000 indicates that the cleaning of the air flow system is needed (for example, fan 123, line 1080 and / or evaporator 110, etc.). In one embodiment, the energy detector 1027 is used to measure the energy provided to the fan motor in the fan 123. If the fan 123 is drawing too much energy or too little energy, then potential airflow problems are indicated (e.g. , ventilation holes blocked or closed, dirty sheets of the fan, dirty evaporator, dirty filter, broken fan band, fan slider, etc.). If the temperature drop across the evaporator 1010 is less than desirable, then the heat removal capacity of the system has been reduced. Such problems can be broadly divided into two categories: refrigerant amount and refrigerant flow rate. If the system 100 has the correct amount of refrigerant charge and the refrigerant is flowing at the desired rate (e.g., as measured by the 1031 and / or 1030 flow detectors), the system must work efficiently and provide the capacity nominal. Problems with the refrigerant amount or flow rate commonly 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 and must be found and repaired. If the system does not work, it is probably an electrical problem that must be found and corrected. If the system 100 starts and runs but does not produce satisfactory cooling, then the amount of heat collected in the evaporator 110 plus the amount of heat from the engine added and the total rejected from the condenser 107 is not the amount of total heat the unit is designed for. to drive. 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 operating temperature 110; (2) Condensation unit condensing temperature and / or (3) Coolant 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 in the unit are the main factors in the classification of efficiency. A larger piping surface results in a lower condensing temperature and a higher EER. A larger evaporation 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 / h by the input watts.

Table 1 Normal operating temperatures of the evaporator 110 can be found by subtracting the division of the design coil from the average air temperature advancing through the evaporator 110. The division of the coil will vary with the design of the system. Systems in the EER range of 7.0 to 8.0 commonly have design divisions in the range of -3.9 ° C (25 ° F) to -1.1 ° C (30 ° F). Systems in the EER range of 8.0 to 9.0 commonly have design divisions in the range of -6.7 ° C (20 ° F) to -3.9 ° C (25 ° F). Systems with EER ratings of 9.0 + will have design divisions in the range of -9.4 ° C (15 ° F) to -6.7 ° C (20 ° F). The formula used to determine the operating temperature of the coil is: - division where COT is the operating temperature of the coil, EAT is the temperature of the air entering the coil (for example, as measured by the temperature detector 1026), LAT is the temperature of the air exiting the coil (for example, such is measured by the temperature detector 1022), and division is the design division temperature. The value (EAT + LAT) 2 is the average air temperature, which is also referred to as the mean temperature difference (BAT). It is also sometimes called the coil TED or? T. "Division" is the design division according to the EER classification. For example, a unit that has an incoming air condition of 26.7 ° C (80 ° F) of dry bulb and a temperature drop of 11.1 ° C (20 ° F) through the evaporator coil 110 will have a coil temperature of determined operation as follows: For an EER classification of 7.0 to 8.0: '80 + 60 ^ COT = - 25a30 ° = 4.4oC (40 ° F) fl7.2oC (45oA) v ¿2- J For an EER rating of 8. 0 to 9. 0 '80 + 60 ^ COT = 20a25 ° = 7.2 ° C (45o) at0 ° C (50 °) l 2 For an EER rating of 9. 0+ Thus, the operating temperature of the coil changes with the EER rating of the unit. The surface area of the condenser 107 affects the condensation temperature that the system 100 must develop to operate at nominal capacity. The variation in the size of capacitor 107 also affects the cost of production and price of the unit. The smaller the capacitor 107, the lower the efficiency rating (EER). In the same EER ratings used for evaporator 110, at 35 ° C (95 ° F) from the external environment, the EER category of 7.8 to 8.0 will operate in the condenser division range of -3.9 ° C (25 ° F) ) at -1.1 ° C (30 ° F), the EER category of 8.0 to 9.0 in the splitting range of condenser 107 from -6.7 ° C (20 ° F) to -3.9 ° C (25 ° F) and the EER category of 9.0+ in the capacitor division interval 107 from -6.7 ° C (20 ° F) to -3.9 ° C (25 ° F) and the EER category of 9.0+ in the division interval of the capacitor 107 -9.4 ° C (15 ° F) to -6.7 ° C (20 ° F). This means that when the air entering condenser 107 is at 35 ° C (95 ° F), the formula for finding the condensing temperature is: RCT = EAT + division where RCT is the condensing temperature of the refrigerant, EAT is the inlet air temperature of the condenser 107 and division is the design temperature difference between the incoming air temperature and the condensing temperatures of both the hot high pressure steam of the compressor 105. For example, using the formula with EAT of 35 ° C (95 ° F), the division for the various EER systems would be: For an EER classification of 7.0 to 8.0 RCT = 95 ° + 25 to 30 ° = 48.9 ° C (120 ° F) to 51.7 ° C (125 °) F) For an EER classification of 8.0 to 9.0 RCT = 95 ° + 20 to 25 ° = 46.1 ° C (115 ° F) to 48.9 ° C (120 ° F) For an EER rating of 9.0+ RCT = 95 ° + 15 to 20 ° = 43.0 ° C (110 ° F) to 46.1 ° C (115 ° F) Operating head pressures vary not only from changes in outdoor temperatures but with different EER classifications. The amount of subcooling produced in the condenser 107 is determined primarily by the amount of refrigerant in the system. 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 has the predominant effect. Therefore, regardless of the EER classifications, the unit must have, if properly charged, a subcooled liquid at -4 ° C (15 ° F) to -6.7 ° C (20 ° F). High ambient temperatures will produce the lowest subcooled liquid due to the amount of refrigerant reduced in the liquid state in the system. More refrigerant will remain in the vapor state to produce higher pressures and condensation temperatures necessary to expel the required amount of heat. Table 1 shows 11 probable causes of problems in the air conditioning system. After each probable cause is the reaction that the cause would have on the low side or suction pressure of the cooling system, the superheat of the evaporator 110, the high side or discharge pressure, the amount of supercooling of the liquid coming out of the condenser 107 and the amperage extracted from the condensing unit. In one embodiment, an air flow detector (not shown) is included to measure the air on the condenser. Insufficient air on the evaporator 110 (as measured, for example, when using the air flow detector 1023 and / or the differential pressure detector 1025) is indicated by a temperature drop greater than that desired in the air at through the evaporator 110. An unbalanced load in the evaporator 110 will also give the opposite indication, indicating that some of the circuits of the evaporator 110 are overloaded while others are slightly charged. In one embodiment, the temperature detector 1022 includes multiple detectors for measuring the temperature through the evaporator. The slightly loaded sections of the evaporator 110 allow the liquid refrigerant to exit the coil and enter the suction manifold and suction line. In TXV systems, the liquid refrigerant passing through the TXV bulb can cause the valve to close. This reduces the operating temperature and capacity of the evaporator 110 also as the suction pressure decreases. The operation superheat of the evaporator 110 may become very low due to the liquid leaving some of the sections of the evaporator 110.

With an inappropriate air flow, the high side pressure or discharge pressure will be low due to the reduced load on the compressor 105, reduced amount of pumped refrigerant vapor and reduced thermal load on the condenser 107. Subcooling of the condenser liquid 107 would be on the high side of the normal range due to the reduction in demand of the refrigerant by the TXV. The unit condensing amperage extracted would be reduced due to the reduced load. In systems using fixed measuring devices, the unbalanced load would produce a lower temperature drop of air through the evaporator 110 because the amount of refrigerant supplied by the fixed measuring device would not be reduced; consequently, the system pressure (boiling point) would be approximately the same. The superheat of the evaporator 110 would fall to zero with the flooding of liquid refrigerant to the suction line. Under an extreme case of imbalance, the liquid returning to the compressor 105 could cause damage to the compressor 105. The reduction in heat accumulated in the evaporator 110 and the decrease in refrigerant vapor to the compressor 105 will decrease the load on the compressor 105. The pressure of compressor discharge 105 (hot gas pressure) will be reduced. The coolant flow rate will only be slightly reduced due to the lower head pressure. The refrigerant subcooling will be in the normal range. The amperage drawn from the condensing unit will be slightly lower due to the reduced load on the compressor 105 and reduction in head pressure. In the case of excessive loading, there is the opposite effect. The temperature drop of the air through the coil will be lower, because the unit can not cool the air as much as it should be. The air moves through the coil at too high a speed. There is also the possibility that the temperature of the air entering the coil is higher than the return air of the conditioned area. This could be from air leaks in the return duct system that draws hot air from unconditioned areas. Excessive loading raises the suction pressure. The refrigerant evaporates at a faster rate than the pumping speed 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 try to maintain the overheating settings. If the system uses fixed measuring devices, the overheating will be high. The fixed measuring devices can not feed enough increase in quantity of refrigerant to keep the evaporator 110 fully active.

The high side pressure or discharge pressure will be high. The compressor 105 will pump more steam due to increase in suction pressure. The condenser 107 should handle more heat and develop a higher condensing temperature to expel the additional heat. A higher condensing temperature means higher pressure on the high side. The amount of liquid in the system has not changed nor is the flow of refrigerant restricted. The liquid subcooling will be in the normal range. The amperage drawn from the unit will be high due to the additional load on the compressor 105. When the temperature of the ambient air entering the condenser 107 is low, then the heat transfer rate of the capacitor 107 is excessive, producing a discharge pressure excessively low. As a result, the suction pressure will be low because the amount of refrigerant through the measuring device will be reduced. This reduction will reduce the amount of liquid refrigerant supplied to the evaporator 110. The coil will produce less steam and the suction pressure drops. The decrease in flow velocity of the refrigerant to the coil reduces the amount of active coil and results in a higher superheat. In addition, the reduced capacity of the system will decrease the amount of heat removed from the air. There will be a higher temperature and higher relative humidity in the conditioned area and the high side pressure will be low. This initiates a reduction in the capacity of the system. The amount of liquid subcooling will be in the normal range. The amount of liquid in the condenser 107 will be higher, but the heat transfer rate of the evaporator 110 is lower. The amperage drawn from the condensing unit will be lower because the compressor 105 is doing less work. The amount of fall in ambient air temperature of the condenser 107 that the air conditioning system will tolerate depends on the type of pressure reducing device in the system. Systems that use fixed measuring devices will have a gradual reduction in capacity as the outside environment drops to 35 ° C (95 ° F). This gradual reduction occurs at 18.3 ° C (65 ° F). At a temperature lower than this temperature the loss of capacity is drastic and some means must be used to maintain the head pressure to prevent the temperature of the evaporator 110 from falling below the freezing point. Some systems control the air through the condenser 107 via dampers in the air stream or a variable speed condenser fan 107. Systems that use TXV will maintain a higher capacity at an ambient temperature of 8.3 ° C (47 ° F). Below this temperature, controls should be used. The control of the air flow through the condenser 107 using dampers or the speed control of the condenser fan 107 can also be used. In larger TXV systems, the amount of liquid in the condenser 107 is used to control the head pressure. The higher the temperature of the air entering the condenser 107, the higher the condensing temperature of the refrigerant vapor to expel the heat in the steam. The higher the condensing temperature, the higher the head pressure. The suction pressure will be high for two reasons: (1) the pump efficiency of the compressor 105 will be lower and (2) the higher liquid temperature will increase the amount of flash gas in the measuring device, further reducing the efficiency of the system. The amount of overheating produced in the coil will be different in a TXV system and a fixed measuring device system. In the TXV system, the valve will keep the superheat close to the limits of its adjustment range although the actual temperatures involved will be higher. In a fixed measuring device system, the amount of superheating produced in the coil is the inverse of the air temperature through the condenser 107. The flow velocity through the fixed measuring devices is directly affected by the head pressure. The higher the air temperature, the higher the head pressure and the higher the flow velocity. As a result of the higher flow velocity, the subcooling is lower. Table 2 shows the overheating that will develop in an appropriately charged air conditioning system using fixed measuring devices. The head pressure will be high at the highest ambient temperatures due to the higher condensing temperatures required. The liquid subcooling of condenser 107 will be in the lower portion of the normal range. The amount of liquid refrigerant in the condenser 107 will be slightly reduced because more will remain in the vapor state to produce the higher pressure and higher condensation temperature. The amperage drawn from the condensing unit will be high. Temperature of air that overheats condenser 107 (° F) (° F) 18.3 ° C (65 ° F) 16.7 ° C (30 ° F) 23.9 ° C (75 ° F) 13.9 ° C (25 ° F) 26.7 ° C (80 ° F) 11.1 ° F (20 ° F) 29.4 ° C (85 ° F) 10 ° C (18 ° F) 32.2 ° C (90 ° F) 8.3 ° C (15 ° F) 35 ° C (95 ° F) 5.5 ° C (10 ° F) 40.5 ° C (105 ° F) and plus 2.8 ° C (5 ° F) Table 2 A shortage or deficit of refrigerant in the system means less liquid refrigerant in the evaporator 110 to collect heat and lower suction pressure. The smaller amount of liquid supplied to the evaporator 110 means less active surface in the coil to vaporize the liquid refrigerant and more surface to raise the temperature of the vapor. Overheating will be high. There will be less steam for the compressor 105 to operate and less head for the capacitor 107 to reject, lower high side pressure and lower condensing temperature. The compressor 105 in an air conditioning system is cooled primarily by the cold return suction gas. The compressor 105 that is under load can have a much higher operating temperature. The amount of subcooling will be less than normal due to none, depending on the amount of underload. The operation of the systems is usually not affected very seriously until the subcooling is zero and the hot gas starts to leave the condenser 107, together with the liquid refrigerant. The amperage drawn from the condensing unit will be slightly less than normal. An overload of refrigerant will affect the system in different ways, depending on the pressure reducing device used in the system and the amount of overload.

In systems that use a TXV, the valve will attempt to control the flow of refrigerant in the coil to maintain the valve's overheating setting. However, the extra refrigerant will return to condenser 107, occupying something of the heat transfer area that would otherwise be available for condensation. As a result, the discharge pressure will be slightly higher than normal, the liquid subcooling will be high and the amperage of the extracted unit will be high. The suction pressure and the superheat of the evaporator 110 will be normal. Excessive overload will cause an even higher head pressure and oscillate the TXV. For TXV system with excessive overload, the suction pressure will be commonly high. Not only does the reduction in capacity of the compressor 105 (due to the higher head pressure) raise the suction pressure, but the higher pressure will cause the TXV valve to be supercharged in its opening stroke. This will cause a wider interval of oscillation of the valve. The superheat of the evaporator 110 will be very erratic from the normal low to liquid range outside the coil. The high side pressure or discharge pressure will be extremely high. The liquid subcooling will also be high due to excessive liquid in the condenser 107. The amperage of the extracted condensing unit will be higher due to the extreme load on the compressor motor 105. The amount of refrigerant in the fixed measuring system has a direct effect on the performance of the system.

An overload has a greater effect than an underload, but both affect system performance, efficiency (EER) and operating cost. Figures 12 to 14 show how the performance of a typical capillary tube air conditioning system is affected by an incorrect amount of refrigerant charge. In Figure 12, at 100% correct load (55 ounces), the unit develops a net capacity of 26,200 Btu / h. When the load amount is varied 5% either in one direction or another, the capacity falls as the load varies. The removal of 5% (3 ounces) of refrigerant reduces the net capacity to 25,000 Btu / h. Another 5% (2.5 ounces) reduces the capacity to 22,000 Btu / h. Hence, the reduction in capacity becomes very drastic: 85% (8 ounces), 18,000 Btu / h; 80% (11 ounces), 13,000 Btu / h; and 75% (14 ounces), 8,000 Btu / h. The overload has a similar effect but at a faster rate of reduction. The addition of 3 ounces of refrigerant (5%) reduces the net capacity to 24,600 Btu / h; 6 ounces added (1059 reduces capacity to 19,000 Btu / h and 8 added ounces (15%) drops capacity to 11,000 Btu / h.) This shows that the overload of one unit has a greater effect per ounce of refrigerant than the Underload Figure 13 is a graph showing the amount of electrical energy the unit demands due to the pressure created by the amount of refrigerant in the system as the refrigerant charge is varied. At 100% load (55 ounces) the unit uses 32 kW. As the load is reduced, the wattage demand also drops, from 29.6 kW to 95% (3 ounces), to 27.6 kW to 90% (6.5 ounces), to 25.7 kW to 85% (8 ounces), to 25 kl to 80% (11 ounces), and 22.4 kW to 75% (14 short ounces of correct load). When the unit is overloaded, the energy consumed also increases. At 3 ounces (5% overload) the energy consumed is 34.2 kW, at 6 ounces (10% overload) 39.5 kW, and at 8 ounces (15% overload), 48 kW. Figure 14 shows the efficiency of the unit (EER classification) based on the capacity in Btu / h of the system against the energy consumed by the condensing unit. At the correct load (55 ounces), the efficiency (EER rating) of the unit is 8.49. As the refrigerant is reduced, the EER rating falls to 8.22 to 9% load, to 7.97 to 90%, to 7.03 to 85%, to 5.2 to 80%, and to 3.57 to 75% full load of the refrigerant. When the refrigerant is added to 5% (3 ounces) the EER rating falls to 7.19. At 10% (6 ounces) the EER is 4.8, and at 15% overload (8 ounces) the EER is 2.29. The overload effect produces a high suction pressure because the flow of the refrigerant to the evaporator 110 increases. The suction superheat decreases due to the additional amount to the evaporator 110. At approximately 8 to 10% overload, the suction superheat becomes zero and the liquid refrigerant will exit the evaporator 110. This causes flooding of the compressor 105 and greatly increases the probability of compressor failures 105. High side pressure or discharge pressure is high due to extra refrigerant in condenser 107. Subcooling of liquid is also high for the same reason. The extracted energy is increased due to the greater amount of steam pumped, also as the highest discharge pressure of the compressor 105. The restrictions in the liquid line 108 reduce the amount of refrigerant to the pressure reducing device 109. Both of the systems TXV valve and fixed measurement device systems will then operate at reduced coolant flow rate to the evaporator 110. The following observations can make the restriction of the liquid line 108. First, the suction pressure will be low due to the reduced amount of refrigerant to the evaporator 110. The suction superheat will be high due to the reduced active portion of the coil, allowing more coil surface to increase the temperature of the steam, also as reduce the boiling point of the refrigerant. The high side pressure or discharge pressure will be low due to the reduced load on the compressor 105. The liquid subcooling will be high. The liquid refrigerant will be accumulated in the condenser 107. It can not flow out at the proper speed due to restriction. As a result, the liquid will cool more than desired. Finally, the amperage drawn from the condensing unit will be low. Either a fixed plugged measuring device or plugged power tube between the TXV valve manifold and the coil will cause part of the coil to be inactive. The system will then operate with an undersized coil, resulting in a low suction pressure because the capacity of the coil has also been reduced. Suction overheating will be high in fixed measuring device systems. The reduced amount of steam produced in the coil and the resulting reduction in suction pressure will reduce the capacity of the compressor 105, head pressure and the flow velocity of the remaining active capillary tubes. The high side pressure or discharge pressure will be low. Subcooling of liquid will be high; the liquid refrigerant will accumulate in the condenser 107. The amperage of the extracted unit will be low. In TXV systems, a plugged feed tube reduces the capacity of the coil. The coil can not provide enough steam to satisfy the pumping capacity of the compressor 105 and the suction pressure is balanced at a low pressure. Overheating, however, will be in the normal range because the valve will adjust to the lowest operating conditions and maintain the set superheat interval. The high side pressure or discharge pressure will be low due to the reduced load on the compressor 105 and condenser 107. The low suction and discharge pressure indicate a refrigerant deficit. Subcooling of liquid is normal to slightly above normal. This indicates an excess 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 drawn from the condensing unit would be low due to the light load on the compressor 105. If the hot gas line 106 is restricted, then the discharge pressure of the high side or of the compressor 105 will be high if it is measured at the output of the compressor 105 or low if it is measured in the outlet line or liquid line of the condenser 107. Either in one case or another, the current drawn from the compressor 105 will be high. The suction pressure is high due to the reduced pumping capacity of the compressor 105. The superheat of the evaporator 110 is high because the suction pressure is high. The high side pressure is high when it is measured at the compressor discharge 105 or low when it is measured in the liquid line. The liquid subcooling is at the high end of the normal range. Even with all this, the amperage extracted from the compressor 105 is above normal. All symptoms point to an extreme restriction in the hot gas line 106. This problem is easily encountered when the discharge pressure is measured at the compressor discharge 105. When the measurement point is the liquid line 108 at the outlet of the capacitor 107, the facts are easily misinterpreted. The high suction pressure and low discharge pressure will usually be interpreted as an inefficient compressor 105. The amperage drawn from the compressor 105 should be measured. The high amperage removed indicates that the compressor 105 is operating against a high discharge pressure. A restriction obviously exists between the output of the compressor 105 and the pressure measuring point. When the compressor 105 will not pump the required amount of refrigerant vapor (eg because it is undersized or is not working at rated capacity). The suction pressure will balance higher than normal. The superheat of the evaporator 110 will be high. The high side pressure or discharge pressure will be extremely low. The liquid subcooling will be low because not much heat will be in the condenser 107. Accordingly, the condensation temperature will be close to the temperature of the incoming air. The amperage drawn from the condensing unit will be extremely low, indicating that the compressor 105 is doing very little work. The following formulas can be used by systems 900, 100 to calculate various operation parameters of the refrigerant cycle system 100 using data from one or more of the detectors shown in FIG. 10. Power is: Watts = vol ts x amps x PF where PF is the power factor. Heat is: Btu = W x? T Specific heat is: Btu = W xcx? T Sensitive heat added or removed from a substance is: Q = W x SH x? T Latent heat added or removed from a substance is: Q = W x LH The cooling effect is: W- - - NRE where W weight of refrigerant passing per minute (eg pounds / minute), 200 Btu / min is the equivalent of one ton of refrigeration and NRE is the net cooling effect (Btu / lb of refrigerant). The coefficient of performance (COP) is: Compression heat cooling effect The capacity of the system is: Qt = 4. 45 x CFM x? H where Qt is the total cooling (sensible and latent) that is made, CFM is the air flow through the evaporator 110 and? H is the change of enthalpy of the air through the coil. The condensing temperature is: RCT = EAT + division where RCT is the condensing temperature of the refrigerant, EAT is the temperature of the air that enters the condenser 107 and division is the difference of design temperature between the temperature of the incoming air and the condensing temperatures of the compressor 105 high-pressure hot steam. The net cooling capacity is. HC = HT - HM where HT is the heat transfer (coarse capacity), HM is the engine heat, HC is the net cooling capacity and PF is the power factor. The air flow velocity of a system can be expressed as: Q = Qs (1.08 x TD) where Q is the flow velocity in CFM, Qs is the sensible heat load in Btu / h and TD is the difference of dry bulb temperature in ° F. In a fan, the air flow (CFM) is roughly related to the rotation (rpm) as follows: CFMz _ rpm2 CFM¡ r? N \ In a fan, the pressure is roughly related to the rotation as follows: In a fan, the work is related to approximately the rotation as follows: In one embodiment, the tachometer 1033 is provided to measure the rotational speed of the fan 123. In one embodiment, the tachometer 1032 is provided to measure the rotational speed of the fan 122. In one embodiment, the system 1000 uses one or more of the equations of previous fans to calculate the desired fan rotation speeds. In one embodiment, the system 1000 controls the speed of the fan 123 and / or the fan 122 to increase the efficiency of the system. The amount of air used for cooling, based on the sensible cooling is approximately: CFM = HS / (TD x 1. 08) The sensible heat removed is: Qs = 1. 08 x CFM x DBT difference The latent heat removed is: Qx = 0. 68 x CFM x g of humidity difference The total heat removed is: o Qt = 4. 5 x CFM x total heat difference The heat transfer rate is: Q = U x A x TD where Q is the heat transfer (Btu / h), U is the global heat transfer coefficient (Btu / h / ft2 / ° F), A is the area (pie2), TD is the temperature difference between the internal and external design temperature and the design temperature of the refrigerated space. The 1050 keyboard is used to provide control inputs to the efficiency verification system. Screen 1008 provides feedback to the user, temperature set point screen. In one embodiment, the use of energy / energy cost may be displayed on the screen 1008. In one embodiment, the system 1000 receives tariff information from the electric utility for use in calculating energy 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 of several detectors in the system 1000 can be displayed on the screen 1008. In one embodiment, diagnostic messages (e.g., change the filter, add refrigerant, etc.) are displayed on the display 1008. In one embodiment, messages of the electric utility are displayed. on the screen 1008. In one embodiment, warning messages of the electricity company are displayed on the 1008 screen., the thermostat 1001 communicates with the electricity company (or other remote device) using power line communication methods such as, for example, BPL. After the system 1000 is configured, the installer programs in the fixed system parameters necessary for the efficiency calculation and / or other quantities derived from the detector data. Typical fixed programmed parameters include type of refrigerant, compressor specifications, condenser specifications, evaporator specifications, duct specifications, fan specifications, system SEER parameters and / or other system parameters . Typical fixed programmed parameters can also include equipment model and / or serial numbers, manufacturer data, design data, etc. In one embodiment, the system 1000 is configured to bring the refrigerant cycle system to the design specifications and then operate the system 1000 in a calibration mode, wherein the system 1000 takes detector readings to measure reference parameters normal for the refrigerant cycle system. Using the measured reference data, system 1000 can calculate various system parameters (eg, split temperatures, etc.). In one embodiment, the system 1000 is first put into operation in a calibration mode to measure reference data and then put into operation in normal verification mode where it compares the operation of the refrigerant cycle system with the reference data. Then the system 100 alerts to potential problems when the operation parameters vary too much from the reference data. In one embodiment, the system 1000 is configured by using a combination of programmed parameters (eg type of refrigerant, temperature divisions, etc.) and reference data obtained by putting the refrigerant cycle system into operation. Figure 15 shows a differential pressure detector 1502 for checking an air filter 1501 in an air handling system. As the filter becomes clogged, the differential pressure across the filter will rise. This increase in differential pressure is measured by the differential pressure detector 1502. The differential pressure measured by the differential pressure sensor 1502 is used to determine the state of the filter 1501. When the differential pressure is too high, then the replacement of the filter 1501. Figure 16 shows differential pressure detector 1502 of Figure 15 provided to a wireless communication unit to allow differential pressure detector data 1502 to be provided to other aspects of verification systems, such as for example the emitter of the condenser unit 1002 or the thermostat 1001. Figure 17 shows the system of Figure 16 implemented using a filter frame 1701 to facilitate retroactive updating of the existing air handling systems. The frame 1701 includes the detector 1502 and the emitter 1601. In the frame 1701 it is configured to fit a standard filter frame. The frame 1701 is configured to retain a standard filter 1501. In one embodiment, the frame 1701 evaluates the cleaning of the filter 1501 by measuring a differential pressure between the air inlet and outlet of the filter. In one embodiment, the frame 1701 evaluates the cleaning of the filter 1501 by providing a light source on the filter side, a light detector on the other side of the filter and by measuring the transmission of light through the filter. In one embodiment, the frame 1701 is calibrated at a reference 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, frame 1701 calibrates a reference light transmission level each time a clean filter is installed. In a modality, frame 1701 indicates that the filter is dirty when the light transmission falls below a percentage of the reference level. Although several modalities have been described above, other modalities will be within the skill of that of ordinary skill in the art. Thus, for example, although it is described primarily in terms of an air conditioning system, that 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, for example, commercial HVAC systems, refrigerator systems, freezers, water coolers, etc. Thus, the invention is limited only by the claims that follow.

Claims

CLAIMS 1. A system for charge control in an electric power system, characterized in that it comprises: a thermostat configured to control a cooling system; a data interface device provided to the thermostat, the data interface device is configured to receive commands or commands, the data interface device is addressable using an identification code, and a remote verification system, the verification system Remote is configured to send a first command or command to the data interface device to adjust the load on the electrical power system.
2. The system according to claim 1, characterized in that the first command comprises a command or shutdown command.
The system according to claim 1, characterized in that the first command comprises a temperature set point command of the thermostat.
The system according to claim 1, characterized in that the first command comprises a command for specifying a temperature set point of the thermostat.
5. The system according to claim 1, characterized in that the first command comprises a command to turn off the cooling system for a specified period of time.
The system according to claim 1, characterized in that the first command comprises a command to reduce a temperature set point for a specified period of time.
The system according to claim 1, characterized in that the first command comprises a command to reduce a temperature set point for a specified period of time.
8. The system according to claim 1, characterized in that the data interface device comprises a modem.
The system according to claim 1, characterized in that the data interface device comprises a broadband modem in an electrical line.
The system according to claim 1, characterized in that the data interface device comprises a wireless modem.
The system according to claim 1, characterized in that the data interface device comprises a telephone modem.
12. A system for charge control in an electric power system, characterized in that it comprises: a cooling system comprising an evaporator unit, - a data interface device provided to the evaporator unit, the interface device of data is configured to receive commands in the electric power system, and a remote verification system, the remote verification system is configured to send a first command to the data interface device to adjust the load in the electric power system .
The system according to claim 12, characterized in that the first command comprises a shutdown command.
The system according to claim 12, characterized in that the first command comprises a command to cause the cooling system to operate in a relatively low energy mode.
The system according to claim 12, characterized in that the first command comprises a command for specifying a temperature set point of the thermostat.
The system according to claim 12, characterized in that the first command comprises a command to turn off the cooling system for a specified period of time.
17. The system according to claim 12, characterized in that the first command comprises a command to reduce a temperature setpoint for a specified period of time.
18. The system in accordance with the claim 12, characterized in that the first command comprises a command to reduce a temperature set point for a specified period of time.
19. The system according to claim 12, characterized in that the data interface device comprises a modem.
The system according to claim 12, characterized in that the data interface device comprises a broadband modem in an electrical line.
21. The system in accordance with the claim 12, characterized in that the data interface device comprises a wireless modem.
22. The system according to claim 12, characterized in that the data interface device comprises a telephone modem.
23. A system for charge control in an electric power system, characterized in that it comprises: a cooling system condenser unit; a compressor provided to the condenser unit; a data interface device provided to the condenser unit, the data interface device is configured to receive commands in the electric power system, and a remote verification system, the remote verification system is configured to send a First command to the data interface device to adjust the load on the electric power system.
24. The system according to claim 23, characterized in that the first command comprises a shutdown command.
25. The system according to claim 23, characterized in that the first command comprises a command to cause the compressor to operate in a mode of relatively low speed.
26. The system in accordance with the claim 23, characterized in that the first command comprises a command to cause the capacitor unit to operate in a relatively low energy mode.
The system according to claim 23, characterized in that the first command comprises a shutdown command to cause the cooling system to turn off for a specified period of time.
The system according to claim 23, characterized in that the first command comprises a command to cause the compressor to operate in a relatively low speed mode for a specified period of time.
29. The system according to claim 23, characterized in that the first command comprises a command to cause the capacitor unit to operate in a relatively low energy mode for a specified period of time.
30. The system according to claim 23, characterized in that the remote verification system is further configured to send a second command to interrogate an operation value of the cooling system.
31. The system according to claim 23, characterized in that the operation value comprises an efficiency value.
32. The system in accordance with the claim 23, characterized in that the data interface device comprises a modem.
33. The system according to claim 23, characterized in that the data interface device comprises a broadband modem in wired line.
34. A system for charge control in an electric power system, characterized in that it comprises: a cooling system comprising: an evaporator unit; a condenser unit; a thermostat; and one or more data interface devices provided to the cooling system, the data interface devices are configured to receive commands; and a remote verification system, the remote verification system is configured to send a first command to the data interface devices to adjust the load on the electric power system.
35. The system according to claim 34, characterized in that the first command comprises a shutdown command.
36. The system according to claim 34, characterized in that the first command comprises a command to cause a compressor in the cooling system to operate in a relatively low speed mode.
37. The system according to claim 34, characterized in that the first command comprises a command to cause the cooling system to operate in a relatively low energy mode.
38. The system in accordance with the claim 34, characterized in that the first command comprises a shutdown command to cause the cooling system to turn off for a specified period of time.
39. The system according to claim 34, characterized in that the first command comprises a command to cause the compressor in the cooling system to operate in a relatively low speed mode for a specified period of time.
40. The system according to claim 34, characterized in that the first command comprises a command to cause the capacitor unit to operate in a relatively low energy mode for a specified period of time.
41. The system according to claim 34, characterized in that the remote verification system is further configured to send a second command to interrogate an operation value of the cooling system.
42. The system according to claim 41, characterized in that the operation value comprises an efficiency value.
43. The system according to claim 34, characterized in that the data interface device comprises a modem.
44. The system according to claim 34, characterized in that the data interface device comprises a broadband modem in an electrical line.
45. A verification system for verifying the operation of a refrigerant cycle system, characterized in that it comprises: a plurality of condenser unit detectors configured to measure operating characteristics of a condenser unit, the plurality of detectors of the refrigerant unit. capacitor comprises a detector for detecting when a compressor in the condenser unit is drawing electrical power, the plurality of detectors of the capacitor unit further comprise at least a first temperature sensor, the capacitor unit comprises a capacitor and a compressor; one or more detectors of the evaporator unit configured to measure one or more operating characteristics of an evaporator unit, the one or more detectors of the evaporator unit comprise at least one second temperature sensor, the evaporator unit comprises a evaporator and an air manipulator fan; one or more environmental detectors configured to measure one or more environmental conditions; and a processing system configured to calculate the efficiency of the refrigerant cycle system using at least a portion of the data of the plurality of detectors of the condenser unit, the one or more detectors of the evaporator unit, and the one or more environmental detectors.
46. The verification system according to claim 45, characterized in that the processing system is configured to calculate the energy usage.
47. The verification system according to claim 45, characterized in that the processing system is configured to calculate the energy costs due to the inefficient operation of the refrigerant cycle system.
48. The verification system according to claim 45, characterized in that the processing system is configured to identify performance problems due to a low air flow.
49. The verification system according to claim 45, characterized in that the processing system is configured to identify performance problems due to excessive loading.
50. The verification system according to claim 45, characterized in that the processing system is configured to identify performance problems due to the underload of the refrigerant.
51. The verification system according to claim 45, characterized in that the processing system is configured to identify performance problems due to the overload of the refrigerant.
52. The verification system according to claim 45, characterized in that the processing system is configured to identify performance problems due to the restriction of the liquid line.
53. The verification system according to claim 45, characterized in that the processing system is configured to identify performance problems due to restriction of the suction line.
54. The verification system according to claim 45, characterized in that the processing system is configured to identify performance problems due to the restriction of the hot gas line.
55. The verification system according to claim 45, characterized in that the processing system is configured to identify performance problems due to the inefficient operation of the compressor.
56. The verification system according to claim 45, characterized in that the processing system is configured to provide data for graphs of energy use and costs.
57. The verification system according to claim 45, characterized in that the processing system is configured to provide data concerning the operation of the refrigerant cycle system to a remote verification center.
58. The verification system according to claim 45, characterized in that the processing system is configured to provide data concerning the operation of the refrigerant cycle system to a remote verification center using power line networks.
59. The verification system according to claim 45, characterized in that the processing system is configured to provide data concerning the operation of the refrigerant cycle system to a remote verification center using broadband in power line networks.
60. The verification system according to claim 45, characterized in that it further comprises an electronically controlled measuring device to allow control of the refrigerant to an evaporator in an energy efficient manner.
61. The verification system according to claim 45, characterized in that the verification system is configured with data with respect to a maximum expected efficiency for the refrigerant cycle system.
62. The verification system according to claim 45, characterized in that the verification system is configured with data with respect to a type of refrigerant used in the refrigerant cycle system.
63. The verification system according to claim 45, characterized in that the verification system is configured with data with respect to the characteristics of the capacitor.
64. The verification system according to claim 45, characterized in that the verification system is configured with data with respect to the characteristics of the evaporator.
65. The verification system according to claim 45, characterized in that the plurality of detectors of the condenser unit comprise a temperature sensor configured to measure the temperature of the refrigerant in a suction line.
66. The verification system according to claim 45, characterized in that the plurality of detectors of the condenser unit comprise a temperature sensor configured to measure the temperature of the refrigerant in a liquid line.
67. The verification system according to claim 45, characterized in that the plurality of detectors of the condenser unit comprise a temperature sensor configured to measure the temperature of the refrigerant in a hot gas line.
68. The verification system according to claim 45, characterized in that the plurality of detectors of the condenser unit comprise a pressure sensor configured to measure the pressure of the refrigerant in a suction line.
69. The verification system according to claim 45, characterized in that the plurality of detectors of the condenser unit comprise a pressure sensor configured to measure the pressure of the refrigerant in a liquid line.
70. The verification system according to claim 45, characterized in that the plurality of detectors of the condenser unit comprise a pressure sensor configured to measure the pressure of the refrigerant in a hot gas line.
71. The verification system according to claim 45, characterized in that the plurality of detectors of the condenser unit comprises at least one flow detector of the refrigerant.
72. The verification system according to claim 45, characterized in that the plurality of detectors of the condenser unit comprises at least one contaminant detector of the refrigerant.
73. The verification system according to claim 45, characterized in that the plurality of detectors of the capacitor unit comprises at least one tachometer of the condenser fan.
74. The verification system according to claim 45, characterized in that the plurality of detectors of the condenser unit comprises at least one temperature sensor configured to measure the temperature of the air outside the condenser.
75. The verification system according to claim 45, characterized in that the at least one of the detectors of the evaporator unit comprises a temperature sensor for measuring the refrigerant temperature to the evaporator.
76. The verification system according to claim 45, characterized in that the at least one of the detectors of the evaporator unit comprises a temperature sensor for measuring the temperature of the refrigerant outside the evaporator.
77. The verification system according to claim 45, characterized in that the at least one of the detectors of the evaporator unit comprises a temperature sensor for measuring the temperature of the air to the evaporator.
78. The verification system according to claim 45, characterized in that the at least one of the detectors of the evaporator unit comprises a temperature sensor for measuring the temperature of the air outside the evaporator.
79. The verification system according to claim 45, characterized in that the at least one of the detectors of the evaporator unit comprises at least one humidity detector.
80. The verification system according to claim 45, characterized in that the at least one of the detectors of the evaporator unit comprises at least one air flow detector.
81. The verification system according to claim 45, characterized in that the at least one of the detectors of the evaporator unit comprises at least one differential pressure detector.
82. A verification system for verifying an evaporator in a refrigerant cycle system, characterized in that it comprises: a first temperature sensor configured to measure the temperature of the inlet air to an evaporator; a second temperature sensor configured to measure the temperature of the evaporator outlet air; one or more environmental detectors configured to measure one or more environmental conditions; a detector to detect when the air is flowing through the evaporator; and a processing system configured to calculate evaporator performance criteria using at least a portion of the data from the first temperature sensor and the second temperature sensor.
83. The verification system according to claim 82, characterized in that the processing system is configured to calculate the efficiency.
84. The verification system according to claim 82, characterized in that the processing system is configured to calculate the energy costs due to the inefficient operation of the evaporator.
85. The verification system according to claim 82, characterized in that the processing system is configured to identify performance problems due to a low air flow.
86. The verification system according to claim 82, characterized in that it also comprises an air flow detector.
87. The verification system according to claim 82, characterized in that it also comprises a tachometer for a fan provided to the evaporator.
88. The verification system according to claim 82, characterized in that it further comprises a third temperature sensor configured to measure the inlet temperature of the evaporator coolant and a third temperature sensor configured to measure the outlet temperature of the evaporator coolant. .
89. The verification system according to claim 82, characterized in that it further comprises one or more pressure detectors configured to measure a pressure differential across the evaporator.
90. The verification system according to claim 82, characterized in that it also comprises at least one humidity detector.
91. The verification system according to claim 82, characterized in that it also comprises one or more electrical detectors for measuring the electric power provided to a fan motor of a fan that supplies air to the evaporator.
92. The verification system according to claim 82, characterized in that it also comprises a coolant flow detector.
93. The verification system according to claim 82, characterized in that the processing system is configured to provide data for graphs of uses and energy costs.
94. The verification system according to claim 82, characterized in that the processing system is configured to provide data concerning the operation of the evaporator to a remote verification center.
95. The verification system according to claim 82, characterized in that the processing system is configured to provide data concerning the operation of the evaporator system to a remote verification center using power line networks.
96. The verification system according to claim 82, characterized in that the processing system is configured to provide data concerning the operation of the evaporator to a remote verification center using brand networks in electric line.
97. The verification system according to claim 82, characterized in that it also comprises an electronically controlled measuring device to allow control of the refrigerant to the evaporator.
98. The verification system according to claim 82, characterized in that the verification system is configured with data with respect to a maximum expected efficiency for the evaporator.
99. The verification system according to claim 82, characterized in that the verification system is configured with data with respect to a type of refrigerant used in the evaporator.
100. The verification system according to claim 82, characterized in that the verification system is configured with data with respect to one or more physical characteristics of the evaporator.
101. The verification system according to claim 82, characterized in that the verification system is configured with data with respect to a cross-sectional area of a conduit provided to the evaporator.
102. The verification system according to claim 82, characterized in that it further comprises a temperature sensor configured to measure the temperature of the refrigerant in a refrigerant line provided to the evaporator.
103. The verification system according to claim 82, characterized in that it also comprises a pressure sensor for measuring the pressure of the refrigerant in an evaporator outlet line.
104. The verification system according to claim 82, characterized in that it also comprises at least one contaminant detector of the refrigerant.
105. The verification system according to claim 82, characterized in that it also comprises a differential pressure detector.
106. A verification system for verifying an evaporator in a refrigerant cycle system, characterized in that it comprises: means for measuring one or more inputs to the evaporator; means for measuring one or more outputs of the evaporator, - programmed data parameters concerned with the operation of the evaporator; and a processing system configured to calculate one or more performance criteria of the evaporator using at least a portion of the data of the means for measuring one or more inputs, the means for measuring one or more outputs and the programmed data parameters, The processing system is configured to provide a performance history of the performance criteria and to calculate an operational efficiency of the evaporator using part of the performance criteria.
107. The verification system according to claim 106, characterized in that the data parameters comprise a type of refrigerant.
108. The verification system according to claim 106, characterized in that the data parameters comprise one or more properties of a refrigerant.
109. The verification system according to claim 106, characterized in that the data parameters comprise one or more calibration values.
110. The verification system according to claim 106, characterized in that the data parameters comprise one or more calibration values obtained from the evaporator during a calibration process.
111. The verification system according to claim 106, characterized in that the data parameters comprise one or more physical properties of the evaporator.
112. The verification system according to claim 106, characterized in that the data parameters comprise one or more dimensional properties of the evaporator.
113. The verification system according to claim 106, characterized in that the one or more entries comprise the temperature of the inlet air.
114. The verification system according to claim 106, characterized in that the one or more inputs comprise the inlet temperature of the refrigerant.
115. The verification system according to claim 106, characterized in that the one or more entrances comprises the electric power provided to a fan of the evaporator.
116. The verification system according to claim 106, characterized in that the one or more inputs comprise the electrical energy provided to a compressor.
117. The verification system according to claim 106, characterized in that the one or more outputs comprise the temperature of the outlet air.
118. The verification system according to claim 106, characterized in that the one or more outputs comprise the outlet coolant temperature.
119. The verification system according to claim 106, characterized in that the one or more outlets comprise the outlet air humidity.
120. The verification system according to claim 106, characterized in that the one or more outlets comprise the air flow.
121. A verification system for verifying a condenser unit in a refrigerant cycle system, characterized in that it comprises: a first temperature sensor configured to measure the refrigerant inlet temperature to a condenser unit; a second temperature sensor configured to measure the coolant outlet temperature of the condenser unit; one or more environmental detectors configured to measure one or more environmental conditions; an electrical detector detecting the energy provided to detect electrical power provided to a compressor of the condenser unit; and a processing system configured to calculate performance criteria of the capacitor using at least a portion of the data from the first temperature detector, the second temperature detector, the environmental detectors and the electrical detector.
122. The verification system according to claim 121, characterized in that the processing system is configured to calculate the efficiency.
123. The verification system according to claim 121, characterized in that the processing system is configured to calculate the energy costs due to the inefficient operation of the condenser unit.
124. The verification system according to claim 121, characterized in that the processing system is configured to identify performance problems due to the underload of the refrigerant.
125. The verification system according to claim 121, characterized in that the processing system is configured to identify performance problems due to the overload of the refrigerant.
126. The verification system according to claim 121, characterized in that it further comprises an air flow detector for a fan provided to the condenser unit.
127. The verification system according to claim 121, characterized in that it further comprises a temperature sensor configured to measure the temperature of the exit air of a condenser coil in the condenser unit.
128. The verification system according to claim 121, characterized in that it further comprises one or more pressure detectors configured to measure a pressure differential of the refrigerant through the compressor.
129. The verification system according to claim 121, characterized in that it also comprises at least one humidity detector.
130. The verification system according to claim 121, characterized in that it further comprises one or more electrical detectors for measuring the electric power provided to a fan motor of the condenser unit.
131. The verification system according to claim 121, characterized in that it also comprises a coolant flow detector.
132. The verification system according to claim 121, characterized in that the processing system is configured to provide data for usage graphs and energy costs.
133. The verification system according to claim 121, characterized in that the processing system is configured to provide data concerning the operation of the capacitor to a remote verification center.
134. The verification system according to claim 121, characterized in that the processing system is configured to provide data concerning the operation of the capacitor system to a remote verification center using electric power networks.
135. The verification system according to claim 121, characterized in that the processing system is configured to provide data concerning the operation of the capacitor to a remote verification center using broadband networks in electric line.
136. The verification system according to claim 121, characterized in that it also comprises a temperature sensor for measuring the temperature of the refrigerant provided to the compressor.
137. The verification system according to claim 121, characterized in that it also comprises a temperature sensor for measuring the temperature of the refrigerant outlet of the compressor.
138. The verification system according to claim 121, characterized in that it further comprises a temperature sensor for measuring the temperature of the coolant outlet of a condenser coil.
139. The verification system according to claim 121, characterized in that the verification system is configured with data with respect to a maximum expected efficiency for the condenser unit under various ambient temperatures.
140. The verification system according to claim 121, characterized in that the verification system is configured with data with respect to a type of refrigerant used in the condenser unit.
141. The verification system according to claim 121, characterized in that it further comprises a pressure detector configured to measure the pressure of the refrigerant in a refrigerant line provided to the compressor.
142. The verification system according to claim 121, characterized in that it further comprises a pressure sensor for measuring the pressure of the refrigerant in an outlet line of the condenser unit.
143. The verification system according to claim 121, characterized in that it also comprises at least one contaminant detector of the refrigerant.
144. The verification system according to claim 121, characterized in that it also comprises an environmental humidity detector.
145. A verification system for verifying a condenser unit in a refrigerant cycle system, characterized in that it comprises: means for measuring one or more inputs to the condenser; means for measuring one or more outputs of the capacitor; programmed data parameters concerning the operation of the capacitor; and a processing system configured to calculate one or more performance criteria of the capacitor using at least a portion of the data of the means for measuring one or more inputs, means for measuring one or more outputs, and programmed data parameters, the system configured to provide a performance history of the performance criteria and calculate an operational efficiency of the capacitor using part of the performance criteria.
146. The verification system according to claim 145, characterized in that the data parameters comprise a type of refrigerant.
147. The verification system according to claim 145, characterized in that the data parameters comprise a property of a refrigerant.
148. The verification system according to claim 145, characterized in that the data parameters comprise one or more calibration values.
149. The verification system according to claim 145, characterized in that it also comprises a pressure sensor for measuring the refrigerant pressure.
150. The verification system according to claim 145, characterized in that the data parameters comprise one or more calibration values obtained from the capacitor during a calibration process.
151. The verification system according to claim 145, characterized in that the data parameters comprise one or more physical properties of the capacitor.
152. The verification system according to claim 145, characterized in that the data parameters comprise one or more dimensional properties of the capacitor.
153. The verification system according to claim 145, characterized in that the one or more entries comprise inlet air temperature.
154. The verification system according to claim 145, characterized in that the one or more entries comprise inlet air temperature.
155. The verification system according to claim 145, characterized in that the one or more inputs comprise electric power provided to the condenser fan.
156. The verification system according to claim 145, characterized in that the one or more inputs comprise electric power provided to the compressor.
157. The verification system according to claim 145, characterized in that the one or more outputs comprise pressure of the output refrigerant.
158. The verification system according to claim 145, characterized in that the one or more outputs comprise a temperature of the outlet coolant.
159. The verification system according to claim 145, characterized in that the one or more outputs comprise coolant flow.
160. An intelligent thermostat to verify the operation of a refrigerant cycle system, characterized in that it comprises: a screen configured to display temperature and efficiency of the system; and a processing system configured to receive detector data from one or more detectors of the condenser unit and one or more detectors from the evaporator unit, the processing system configured to calculate an efficiency of the refrigerant cycle system using minus a portion of the detector data and to display a parameter related to efficiency.
161. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to calculate and display power usage of the HVAC system.
162. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to calculate and display an energy cost of the use of the HVAC system.
163. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to diagnose performance problems.
164. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to diagnose performance problems due to excessive loading.
165. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to diagnose performance problems due to underloading of the refrigerant.
166. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to diagnose performance problems due to refrigerant overload.
167. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to diagnose performance problems due to the restriction of the liquid line.
168. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to identify performance problems due to restriction of the suction line.
169. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to identify performance problems due to the restriction of the hot gas line.
170. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to identify performance problems due to the inefficient operation of the compressor.
171. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to provide data for usage graphs and energy costs.
172. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to provide data concerning the operation of the refrigerant cycle system to a remote verification center.
173. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to provide data concerning the operation of the refrigerant cycle system to a remote verification center using an electric power network.
174. The intelligent thermostat according to claim 60, characterized in that the processing system is configured to provide data concerning the operation of the refrigerant cycle system to a remote verification center using broadband networks in electric line.
175. The intelligent thermostat according to claim 60, characterized in that it further comprises an electronically controlled measuring device to allow the control of refrigerant to an evaporator in an energy efficient manner.
176. The intelligent thermostat according to claim 60, characterized in that the intelligent thermostat is configured with data with respect to a maximum expected efficiency for the refrigerant cycle system.
177. The intelligent thermostat according to claim 60, characterized in that the intelligent thermostat is configured with data with respect to the type of refrigerant used in the refrigerant cycle system.
178. The intelligent thermostat according to claim 60, characterized in that the intelligent thermostat is configured with data with respect to the characteristics of the capacitor.
179. The intelligent thermostat according to claim 60, characterized in that the intelligent thermostat is configured with data with respect to the characteristics of the evaporator.
180. The intelligent thermostat according to claim 60, characterized in that the detectors of the condenser unit comprise a temperature sensor configured to measure the temperature of the refrigerant in a suction line.
181. The intelligent thermostat according to claim 60, characterized in that the detectors of the condenser unit comprise a temperature sensor configured to measure the temperature of the refrigerant in a liquid line.
182. The intelligent thermostat according to claim 60, characterized in that the detectors of the condenser unit comprise a temperature sensor configured to measure the temperature of the refrigerant in a hot gas line.
183. The intelligent thermostat according to claim 60, characterized in that the detectors of the condenser unit comprise a pressure sensor configured to measure the pressure of the refrigerant in a suction line.
184. The intelligent thermostat according to claim 60, characterized in that the detectors of the condenser unit comprise a pressure sensor configured to measure the pressure of the refrigerant in a liquid line.
185. The intelligent thermostat according to claim 60, characterized in that the detectors of the condenser unit comprise a pressure sensor configured to measure the pressure of the refrigerant in a hot gas line.
186. The intelligent thermostat according to claim 60, characterized in that the detectors of the condenser unit comprise at least one coolant flow detector.
187. The intelligent thermostat according to claim 60, characterized in that the detectors of the condenser unit comprise at least one contaminant detector of the refrigerant.
188. The intelligent thermostat according to claim 60, characterized in that the detectors of the condenser unit comprise at least one tachometer of the condenser fan.
189. The intelligent thermostat according to claim 60, characterized in that the detectors of the condenser unit comprise at least one temperature sensor configured to measure the temperature of the air outside the condenser.
190. The intelligent thermostat according to claim 60, characterized in that the detectors of the evaporator unit comprise a temperature sensor for measuring the refrigerant temperature in the evaporator.
191. The intelligent thermostat according to claim 60, characterized in that the detectors of the evaporator unit comprise a temperature sensor for measuring the coolant temperature outside the evaporator.
192. The intelligent thermostat according to claim 60, characterized in that the detectors of the evaporator unit comprise a temperature sensor for measuring the temperature of the air in the evaporator.
193. The intelligent thermostat according to claim 60, characterized in that the detectors of the evaporator unit comprise a temperature sensor for measuring the temperature of the air outside the evaporator.
194. The intelligent thermostat according to claim 60, characterized in that the detectors of the evaporator unit comprise at least one humidity detector.
195. The intelligent thermostat according to claim 60, characterized in that the detectors of the evaporator unit comprise at least one air flow detector.
196. The intelligent thermostat according to claim 60, characterized in that the detectors of the evaporator unit comprise at least one differential pressure detector.
197. The intelligent thermostat according to claim 60, characterized in that it also comprises a modem, the intelligent thermostat configured to report performance criteria to a verification system using said modem.
198. The intelligent thermostat according to claim 60, characterized in that it further comprises a modem, the intelligent thermostat configured to receive shutdown instructions using said modem.
199. The intelligent thermostat according to claim 60, characterized in that it also comprises a modem, the intelligent thermostat configured to receive commands using said modem.
200. The intelligent thermostat according to claim 60, characterized in that it further comprises a modem, the intelligent thermostat configured to receive operating instructions using said modem.
201. A verification system for verifying an air filter in a forced air heating or cooling system, characterized in that it comprises: a differential pressure detector configured to measure a pressure drop across a filter element; and a processing system configured to calculate a filter performance criterion using at least a portion of the differential pressure detector data.
202. The verification system according to claim 201, characterized in that the processing system is configured to indicate when the filter needs to be replaced.
203. The verification system according to claim 201, characterized in that the processing system is configured to calculate energy costs due to the inefficient operation of the filter.
204. The verification system according to claim 201, characterized in that the processing system is configured to identify performance problems due to a low air flow.
205. The verification system according to claim 201, characterized in that it also comprises an air flow detector.
206. The verification system according to claim 201, characterized in that it further comprises a wireless transmitting system for transmitting data from the differential pressure detector to the processing system.
207. The verification system according to claim 201, characterized in that the differential pressure detector is provided to a frame configured to retain a conventional filter element, the frame configured to be mounted on a conventional filter element carrier.
208. The verification system according to claim 201, characterized in that it also comprises a timer, the processor configured to indicate replacement of the filter when the timer exceeds the specific time of use for the filter or when a pressure drop through the filter exceeds a specified amount.
209. A verification system for verifying an air filter in a forced air heating or cooling system, characterized in that it comprises: means for retaining a filter element; means for measuring a pressure drop across the filter element; means to send data to a processing system; and a processing system configured to calculate one or more filter performance criteria using media data to measure a pressure drop.
210. A verification system for verifying an air filter in a forced air heating or cooling system, characterized in that it comprises: means for retaining a filter element; means for measuring light transmission through the filter element; means to send data to a processing system; and a processing system configured to calculate one or more filter performance criteria using the data of the means for measuring light transmission, the processing system configured to establish a reference light transmission value when the filter element is replaced and to indicate replacement of the filter when the light transmission value falls below a threshold value in relation to the reference light transmission value.
211. A verification system for verifying an air filter in a forced air heating or cooling system, characterized in that it comprises: a light source configured to illuminate a portion of the filter element; a light detector configured to receive light from a light source that has passed through the filter element; and a processing system configured to calculate a filter performance criterion using at least a portion of the data of the light source, the processing system configured to establish a reference light transmission value when the filter element is newly installed and to indicate replacement of the filter when the light transmission value falls below a threshold value with respect to the reference light transmission value.
212. The verification system according to claim 201, characterized in that the processing system is configured to identify performance problems due to a low air flow caused by a dirty filter element.
213. The verification system according to claim 201, characterized in that it also comprises an air flow detector.
214. The verification system according to claim 201, characterized in that it further comprises a wireless transmitting system for transmitting data to an HVAC verification system.
215. The verification system according to claim 201, characterized in that the light source is provided to a frame configured to retain the filter element, the frame configured to be mounted on a conventional filter element holder.
216. The verification system according to claim 201, characterized in that it further comprises a timer, a processor configured to indicate replacement of the filter when the timer exceeds the specified usage time for the filter element or when a transmission of light through of the filter falls below a specified amount. SUMMARY OF THE INVENTION A real-time verification system that verifies several aspects of the operation of a refrigerant cycle system is described. In one embodiment, the system includes a processor that measures the energy provided to the refrigerant cycle system and collects data from one or more detectors and uses the detector data to calculate a merit number related to the efficiency of the system. In one embodiment, the detectors include one or more of the following detectors: a suction line temperature detector, a suction line pressure detector, a suction line flow detector, a suction line detector, the temperature of the hot gas line, a pressure detector of the hot gas line, a flow detector of the hot gas line, a temperature detector of the liquid line, a detector of the pressure of the line of the liquid, a detector of the flow of the liquid line. In one embodiment, the detectors include one or more of an evaporator air temperature input detector, an evaporator air temperature output detector, an evaporator air flow detector, an air humidity detector of the evaporator and a differential pressure detector. In one embodiment, the detectors include one or more of a condenser air temperature input detector, a condenser air temperature output detector and a condenser air flow detector, an air humidity detector of the evaporator. In one embodiment, the detectors include one or more of an ambient air detector and an ambient humidity detector.