Patent Publication Number: US-2015059367-A1

Title: Active charge control methods for vapor cycle refrigeration or heat pump systems

Description:
CROSS-REFERENCE TO RELATED APPLICATIONS 
     This application claims the benefit of priority under 35 U.S.C. §119(e) to U.S. Provisional Application Ser. No. 61/873,547, filed Sep. 4, 2013, incorporated herein by reference in its entirety. 
    
    
     STATEMENT REGARDING FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT 
     This invention was made with government support under contract FA8650-04-2403-0017, awarded by The Air Force Research Laboratories (AFRL). The government has certain rights in the invention. 
    
    
     TECHNICAL FIELD 
     This application relates to vapor compression systems and control methods for the vapor compression systems and, more particularly, to vapor compression systems that include means for active charge control and to active charge-control methods for such vapor compression systems. 
     BACKGROUND 
     Vapor compression systems are ubiquitous and have been used for over 100 years to accomplish thermal management in many contexts. Vapor compression systems are found in a small scale in devices such as home refrigerators and air conditioners and in a larger scale in industrial HVAC systems. Small vehicles such as automobiles use vapor compression systems for air conditioning, and larger vehicles such as refrigeration tractor trailers, railroad refrigeration cars, and airline transport and cargo jets use vapor compression systems to control temperature in confined volumes. In most of these applications, the loads and sinks vary relatively slowly, primarily by diurnal cycling. As a specific example, a building sinking temperature will vary throughout the diurnal cycle and with seasonal changes. Nearly all of these changes occur on the scale of hours, thus appearing relatively steady-state. 
     The application of vapor compression systems has been limited when the vapor compression systems are located where drastic changes in load or sink temperatures are present, such as in aviation systems. The challenge to the aviation systems designer, for example, is to accommodate not only large dynamic swings in load but also major swings in both the sink temperature and the sinking source. Current systems in aviation have changes that occur in minutes rather than hours, and future systems are envisioned to have changes that occur in seconds rather than hours. As such, there are ongoing needs for energy-efficient methods for controlling vapor compression systems, small and large, which experience large dynamic swings not only in load but also major swings in both the sink temperature and the sinking source. 
     SUMMARY 
     According to some embodiments, a vapor compression system includes a refrigeration loop, a charge control loop, a high-side sensor, a sink sensor, and a control apparatus. The refrigeration loop configured to transfer heat from a load location to a rejection apparatus at a rejection location. The refrigeration loop includes a plurality of components in fluidic communication through refrigeration lines containing a refrigerant. The plurality of components includes a compressor, a condenser, an expansion valve, and an evaporator at the load location. The charge control loop having an extraction connection and an injection connection to the refrigeration loop. Both the extraction connection and the injection connection place the charge control loop in fluidic communication with the refrigeration loop between the condenser and the expansion valve. The charge control loop has a plurality of components including a discharge control valve, a transfer pump, and a mass storage vessel between the discharge control valve and the transfer pump. The discharge control valve is controllably configured to allow refrigerant to pass from the refrigeration loop through the extraction connection and into the mass storage vessel. Additionally, the transfer pump is controllably configured to inject refrigerant from the mass storage vessel into the refrigeration loop through the injection connection. The high-side sensor is configured to measure a high-side pressure and compressor discharge temperature of a high-pressure side of the refrigeration loop. The sink sensor is configured to measure a sink temperature of the refrigeration loop. The control apparatus electronically is coupled to the compressor, the at least one expansion valve, the discharge control valve, the transfer pump, the high-side sensor, and the sink sensor. 
     According to some embodiments, methods for controlling a vapor compression system include operation of a vapor compression system. The vapor compression system may include a refrigeration loop configured to transfer heat from a load location to a rejection apparatus at a rejection location. The refrigeration loop includes a plurality of components in fluidic communication through refrigeration lines containing a refrigerant. The plurality of components includes a compressor, a condenser, an expansion valve, and an evaporator at the load location. The vapor compression system also includes a charge control loop having an extraction connection and an injection connection to the refrigeration loop, both the extraction connection and the injection connection placing the charge control loop in fluidic communication with the refrigeration loop between the condenser and the expansion valve. The charge control loop includes a plurality of components. The plurality of components in the charge control loop includes a discharge control valve, a transfer pump, and a mass storage vessel between the discharge control valve and the transfer pump. The discharge control valve is controllably configured to allow refrigerant to pass from the refrigeration loop through the extraction connection and into the mass storage vessel. The transfer pump is controllably configured to inject refrigerant from the mass storage vessel into the refrigeration loop through the injection connection. The vapor compression system also includes a high-side sensor that measures a high-side pressure and compressor discharge temperature of a high-pressure side of the refrigeration loop. Further, the vapor compression system includes a sink sensor that measures a sink temperature of the refrigeration loop. Additionally, the vapor compression system also includes a control apparatus electronically coupled to the compressor, the at least one expansion valve, the discharge control valve, the transfer pump, the high-side sensor, and the sink sensor. Thus, the methods for controlling the vapor compression system may include selecting a saturated discharge temperature set-point and selecting an approach high set-point. The methods may also include operating the vapor compression system to transfer heat from the load location to the rejection location. Further, the methods may include polling the sink temperature from the sink sensor and polling the high-side pressure and compressor discharge temperature from the high-side sensor. Additionally, the methods may include determining an approach as a temperature difference between the sink temperature polled from the sink sensor and the compressor discharge temperature polled from the high-side sensor as well as determining a saturated-discharge temperature, the saturated discharge temperature being determined from the high-side pressure polled from the high-side sensor. The method further includes adjusting continually with the control apparatus, while the vapor compression system is operating, one or more of the discharge control valve in response to the temperature difference so as to adjust flow of refrigerant to from the refrigeration loop through the extraction connection and into the mass storage vessel and the transfer pump in response to the saturated-discharge temperature differential so as to adjust flow of refrigerant from the mass storage vessel into the refrigeration loop through the injection connection. 
     Additional features and advantages of the embodiments described herein will be set forth in the detailed description which follows, and in part will be readily apparent to those skilled in the art from that description or recognized by practicing the embodiments described herein, including the detailed description which follows, the claims, as well as the appended drawings. 
     It is to be understood that both the foregoing general description and the following detailed description describe various embodiments and are intended to provide an overview or framework for understanding the nature and character of the claimed subject matter. The accompanying drawings are included to provide a further understanding of the various embodiments, and are incorporated into and constitute a part of this specification. The drawings illustrate the various embodiments described herein, and together with the description serve to explain the principles and operations of the claimed subject matter. 
    
    
     
       BRIEF DESCRIPTION OF THE DRAWINGS 
         FIG. 1  is a schematic diagram of a conventional vapor compression system; 
         FIG. 2  is an illustrative refrigerant pressure-enthalpy graph; 
         FIG. 3  is an illustrative refrigerant pressure-enthalpy graph, plotting conditions used in a conventional vapor compression system during operation; 
         FIG. 4  is a schematic diagram of an exemplary vapor compression system configured according to embodiments described herein; and 
         FIG. 5  is a schematic diagram of an exemplary multiple-load vapor compression system configured according to embodiments described herein. 
         FIG. 6  is a flow chart of an exemplary active charge control scheme according to embodiments described herein. 
     
    
    
     DETAILED DESCRIPTION 
     The vapor compression system and methods for controlling the vapor compression system and for controlling a multiple-load vapor compression system according to various embodiments herein may provide substantial improvements over conventional vapor compression systems, particularly with regard to energy efficiency. To provide a contextual basis for the embodiments of the system and methods, a conventional vapor compression system and its typical control protocol will first be described with reference to FIGS. 1- 3 . Embodiments of a vapor compression system and methods for controlling the vapor compression system will be described below with reference to  FIG. 4 . Embodiments of methods for controlling a multiple-load vapor compression system will be described in detail below with reference to  FIG. 5 . 
     A conventional vapor compression system  1  is shown schematically in  FIG. 1  and is described in detail in U.S. application Ser. No. 14/153,471 which is incorporated herein by reference in its entirety. The conventional vapor compression system  1  typically includes a refrigeration loop  10  that transfers heat from a load location  50  to a rejection location  70 . The heat is transferred via refrigeration lines containing a refrigerant. As the refrigerant travels through the refrigeration loop  10 , it undergoes multiple phase changes effected by various apparatus in fluidic communication with the refrigeration loop  10 . For example, a phase change from liquid to vapor absorbs heat from the load location  50 , and a phase change from vapor back to liquid results in heat being exhausted to the external surroundings at the rejection location  70 . 
     Aspects of the conventional vapor compression system  1  will now be described conceptually and by way of generality, without any intention to capture every nuance and theoretical concept involved in the conventional system. In general, in the conventional vapor compression system  1 , refrigerant vapor on a low-pressure side  12  of the refrigeration loop  10  is compressed by the compressor  20  into a high-pressure side  14  of the refrigeration loop  10 . At least one low-side sensor  37  is located between the evaporator  40  and the compressor  20 . The at least one low-side sensor  37  may measure temperature, pressure, or both. Before entering the compressor  20 , the refrigerant vapor is at a compressor suction temperature (T S ) and a suction pressure (P S ), one or both of which being measurable by the low-side sensor  37 . A saturated suction temperature (T SS ) may be determined from the suction pressure (P S ) using a pressure-enthalpy table or other suitable table for the refrigerant involved. An amount of superheating present in the system is equal to T S −T SS . Compression of the refrigerant by the compressor  20  raises the temperature of the refrigerant proportionally to the amount of compression, i.e., the work performed on the refrigerant to cause the compression. For heat to be removed from the compressed refrigerant vapor, the temperature of the compressed refrigerant vapor must be greater than the temperature of the external surroundings at the rejection location  70 . For example, if the external surroundings at the rejection location  70  are 35° C. on a hot summer&#39;s day, the compressed refrigerant vapor must have a temperature greater than 35° C. for the necessary heat exchange to occur. 
     Heat is removed from the refrigerant at the condenser  60 . The condenser  60  provides communication between the refrigeration loop  10  and the rejection location  70  and eventually the external surroundings. The external surroundings may be the environment, for example, or may be a load location to an additional refrigeration loop that further cools a thermal transfer medium such as air or liquid containing the heat that is rejected from the refrigerant. The refrigeration loop  10  in the condenser is in thermal communication with a rejection apparatus  65 . The rejection apparatus  65  places the condenser  60  and refrigeration loop  10  in thermal communication with the rejection location  70 . 
     The rejection apparatus  65  may be any apparatus that continually rejects heat to the rejection location  70  or the external environment. For example, the rejection apparatus  65  may be a heat sink, a fluid cooling loop, an exhaust fan, or a vent. The rejection apparatus may circulate or otherwise facilitate thermal transfer to a liquid or gaseous coolant medium such as air, water, or glycol that is colder than the refrigerant in the condenser  60 . A sufficient amount of heat is removed from the refrigerant at the condenser  60  to cause the compressed refrigerant vapor to condense to a compressed liquid, generally a saturated liquid or a sub-cooled liquid. A high-side sensor  80  between the condenser  60  and a metering device  30  may measure temperature of this liquid, its pressure, or both its temperature and pressure. The temperature of the compressed liquid at this stage, as measureable by the high-side sensor  80  is called the condenser discharge temperature (T CD ). The pressure of the compressed refrigerant at this stage is s discharge pressure (P D ). A saturated discharge temperature (T SD ) may be determined from the discharge pressure (P D ) using a pressure-enthalpy table or other suitable table for the refrigerant involved. An amount of subcooling in the system is equal to T SD −T CD . 
     Compressed liquid refrigerant re-enters the low-pressure side  12  of the refrigeration loop  10  from the high-pressure side  14  at the metering device  30 . Many types of metering devices are known. Thermal expansion valves are common examples of metering devices. The metering device  30  allows at least a portion of the compressed liquid refrigerant to expand adiabatically into the low-pressure side  12 . This expansion will chill the liquid refrigerant. Depending on the thermodynamic characteristics of the refrigerant, after the expansion the refrigerant may remain in the liquid phase or may consist of some liquid-phase refrigerant and some vapor-phase refrigerant. 
     The chilled refrigerant then travels to the evaporator  40 . At the evaporator  40 , the refrigeration loop  10  is in thermal communication with an evaporator loop  45  containing a circulating thermal transfer medium such as a liquid or gas. The thermal transfer medium transfers heat from the load location  50 . The load location  50  is at a temperature that defines an evaporator load of the vapor compression system. The refrigeration loop  10 , absorb heat from the load location  50 , which has the effect of cooling the space in the vicinity of the load location  50 . The amount of heat being transferred from the load location  50  may be assessed from measuring the output temperature T L  with a load sensor  42 , which may be an output temperature sensor, for example. As illustrative examples, the load location  50  may be a room being air conditioned or the inside of a refrigerator. Fans or other apparatus not shown in  FIG. 1  may be used to increase the transfer of heat at the evaporator  40 . As the heat is transferred from the load location  50  to the refrigerant at the evaporator  40 , the refrigerant evaporates to form a refrigerant vapor. This refrigerant vapor then is directed back to the compressor  20  and continues around the refrigeration loop  10  again. 
     Control, output, and energy efficiency of the conventional vapor compression system  1  are constrained by the mechanical characteristics of the components generally present in the system. For example, typical compressors used for compressing the vapor refrigerant may malfunction or even be rendered inoperative if any liquid-phase refrigerant enters the intake of the compressor  20 . As such, the conventional vapor compression system  1  typically is operated such that the refrigerant vapor is superheated at the evaporator  40  by a certain margin such as 5° C., for example, thereby ensuring that no liquid-phase refrigerant is present at the intake of the compressor  20 . In the conventional vapor compression system  1  of  FIG. 1 , superheating of the refrigerant vapor may be ensured by the combination of the evaporator  40  and a metering device  30  that affects the amount of compressed liquid refrigerant entering the low-pressure side  12  of the refrigeration loop  10  through the metering device  30  to the heat load in the evaporator  40 . 
     Typically, the metering device  30  is designed to adjust the refrigerant flow based on the measured temperature difference between the low-side sensor  37  and the metering device sensor  32 . An exemplary metering device  30  is an expansion valve. To illustrate, the metering device  30  may be set to maintain a threshold difference such as5° C. Thereby, when the measured temperature difference between the low-side sensor  37  and the metering device sensor  32  is greater than5° C., the metering device  30  may open more widely to allow more refrigerant to enter the low-pressure side  12 . A temperature difference of greater than5° C. in this illustration would indicate that the evaporator  40  can evaporate a greater amount of liquid refrigerant while ensuring no liquid enters the compressor  20 . Conversely, when the measured temperature difference between the low-side sensor  37  and the metering device sensor  32  is less than5° C., the metering device  30  may open less widely to allow less refrigerant to enter the low-pressure side  12 . A temperature difference of less than5° C. in this illustration would indicate that more refrigerant is entering the evaporator  40  than the evaporator  40  has capacity to evaporate while still ensuring no liquid enters the compressor  20 . 
     Thermodynamic aspects of the conventional vapor compression system  1  are illustrated through the pressure-enthalpy graphs of  FIGS. 2 and 3 .  FIG. 2  shows a generic pressure-enthalpy graph  100  of a refrigerant. Pressure (P) is plotted as a function of enthalpy (H). The curve in the generic pressure-enthalpy graph  100  includes a liquid line  110  and a vapor line  120  that meet at a critical point  130 . To the left (lower enthalpy side) of the liquid line  110  is a liquid zone  115 , in which the refrigerant is exclusively in the liquid phase. To the right (higher enthalpy side) of the vapor line  120  is a vapor zone  125 , in which the refrigerant is exclusively in the vapor phase. At the liquid line  110 , the refrigerant is a saturated liquid. In a two-phase zone  140  between the liquid line  110  and the vapor line  120 , the refrigerant is a two-phase mixture of liquid refrigerant and vapor refrigerant. At the vapor line  120 , the refrigerant is a saturated vapor. At a given pressure, the horizontal distance between the liquid line  110  and the vapor line  120  is the heat of vaporization  145  of the refrigerant at the given pressure, which reflects the amount of energy (i.e., latent heat) that must be added to the refrigerant to change a saturated liquid into a saturated vapor or the amount of energy (i.e., latent heat) that must be removed from the refrigerant to change a saturated vapor into a saturated liquid. 
     A first isotherm  150  and a second isotherm  155  illustrate two lines of constant temperature in the generic pressure-enthalpy graph  100 . The first isotherm  150  represents a higher temperature than the second isotherm  155 . The vertical portions of the isotherms  150 ,  155  to the left of the liquid line  110  illustrate that single-phase liquid refrigerants expand adiabatically, or at a constant enthalpy, until saturation is reached and the liquid refrigerant begins to vaporize. The vertical portions of the isotherms  150 ,  155  to the left of the liquid line  110  also illustrate that removal of heat, moving from the first isotherm  150  to the second isotherm  155 , from single-phase liquid refrigerant at a constant pressure reduces the temperature of the refrigerant. The horizontal portions of the isotherms  150 ,  155  in the two-phase zone  140  illustrate that the process of vaporization or condensation occurs at a constant temperature if constant pressure is maintained. The horizontal portions also illustrate that temperature decreases if pressure is decreased. The tail portions of the isotherms  150 ,  155  to the right of the vapor line  120  illustrate that adding heat to a single-phase vapor refrigerant while maintaining constant pressure will result in an increase of the temperature of the single-phase vapor refrigerant. Or in reverse, the tail portions of the isotherms  150 ,  155  to the right of the vapor line  120  illustrate that removing heat from a single-phase vapor refrigerant while maintaining constant pressure will result in a decrease of the temperature of the single-phase vapor refrigerant until liquid begins to condense at the vapor line  120 . Whereas removal of heat from a single-phase liquid or a single-phase vapor results in a temperature change (from loss or gain of sensible heat), the processes of vaporization and condensation involve addition or removal of latent heat, the heat required to complete a phase change, at a constant temperature and constant pressure. 
     In  FIG. 2 , two lines of constant entropy  157 ,  159  are provided as illustrations. During compression of a single-phase vapor refrigerant, the compression ideally occurs along a line of constant entropy or along a pathway that at least results in an increase of enthalpy. As is clear from  FIG. 2 , the lines of constant entropy  157 ,  159  both cross the first isotherm  150  and the second isotherm  155 . Thus, the compression of vapor-phase refrigerant results in an increase of the temperature of the refrigerant. 
     The operational schematic  101  of  FIG. 3  illustrates how the pressure-enthalpy relationship of a refrigerant is used in the conventional vapor compression system. The pressure-enthalpy curve  105  has a shape typical for common refrigerants used in the art of vapor compression systems but can vary slightly based on the actual refrigerant with no change to the principles involved. At pre-compression point  160 , the refrigerant is a single-phase vapor at suction pressure P S  and is superheated above a saturated suction temperature T SS  by a superheat margin  187  to ensure no liquid refrigerant enters the compressor. Saturated suction conditions are defined by the properties of the specific refrigerant. The saturated suction temperature (T SS ) is defined as the temperature of the refrigerant that is inside the dome at the compressor inlet suction pressure (P S ). In  FIG. 3 , points  180  and  185  are both practically at P S . Saturated discharge temperature (T SD ) is defined as the refrigerant temperature inside the dome at the pressure (P SD ). In  FIG. 3 , point  170  is at P SD  and T SD . Additionally, superheat may be defined as the temperature of the vapor above the saturation temperature, the temperature above the right side of the dome. In  FIG. 3 , points  160  and  165  are both superheated. Likewise, subcooling may be defined as the temperature below the saturated temperature, or points to the left of the dome. In  FIG. 3 , only point  175  is subcooled. The vapor-phase refrigerant is compressed by the compressor to post-compression point  165  at discharge pressure P D . The line from the pre-compression point  160  to the post-compression point  165  nearly follows a path of constant entropy, but the path up the line of constant entropy crosses several isotherms (not shown, see  FIG. 2 ) and thereby results in an increase of temperature. The pressure increase  190  (P D −P S ) involved with this process is directly related to the amount of work performed on the refrigerant and, thereby, to the amount of energy needed to power the compressor during operation of the system. The work is the change in the product of the change in enthalpy from points  160  to  165  and the mass of refrigerant that experienced this change. 
     At the condenser, heat is removed from the vapor-phase refrigerant to cause the refrigerant to condense at a constant saturated discharge temperature (T SD ) that is characteristic of two-phase refrigerant at pressure P D . The refrigerant condenses to saturation point  170  and then typically is subcooled below T SD  by a subcooling margin  172  to reach a subcooled liquid point  175 . Subcooling is not necessarily required but is beneficial for increasing refrigeration system overall cooling capacity and increasing the overall process efficiency. The subcooled liquid is then allowed to expand adiabatically (at constant enthalpy) through the metering device  30  to expanded two-phase point  180 . This expansion occurs at a constant temperature until the liquid begins to vaporize, at which point further expansion results in a decrease of temperature. Then, at the evaporator the remaining liquid refrigerant vaporizes at a constant temperature to saturated vapor point  185 . As noted above, additional heat is typically added to the saturated vapor to produce the superheat margin  187  until the refrigeration cycle again reaches the pre-compression point  160 . 
     Though the operational principles of the conventional vapor compression system  1  illustrated above have been used successfully for years, significant inefficiency is inherent in the conventional vapor compression system  1 . Typically in the conventional vapor compression system  1  of  FIG. 1 , a desired set-point for the output temperature T L  measured at the load sensor  42  at the evaporator load may be chosen. System capacity, i.e., the amount of heat being removed from the load location  50 , then generally is modulated by either cycling the compressor  20  off and on or adjusting the speed of the compressor  20  based on feedback from the load sensor  42 . Modulation of system capacity in this manner is relatively slow dynamically, and overcompensation or undercompensation of adjustments to the cycle of the compressor  20  may result in wasted energy while the desired set-point temperature is achieved. 
     Typically in the conventional vapor compression system  1  of  FIG. 1 , the sinking or rejection temperature can change over a period of time along a range of temperatures. An example could be a building refrigeration or central air system in a high altitude desert location wherein the temperature of the outside air could easily swing 40° F. (22° C.) or more within an average day and the mean daily temperature may swing 60° F. (33° C.) over the seasonal changes in a year. The average house located in the central portions of the United States can have more modest sinking temperature swings yet still experience changes over the seasons. Conversely, an aircraft can have much larger variations in sinking temperature, from 120° F. (49° C.) to −50° F. (−46° C.) in transition from a sunny tarmac to cruising altitude. In all vapor compression systems, the cycle efficiency and required compressor power varies as a function of the system charge and the difference between the sinking temperature and the P SD /T SD , with the limit being the difference between the sinking temperature and the T CD . Ideally, one wants to operate the system at the lowest P SD /T SD  and the lowest T CD . Given a P SD /T SD  and a sinking temperature the system charge directly impacts the systems T CD . The lower the charge, the higher the T CD . Therefore it is preferred to increase the charge as the sinking temperature reduces. 
     In the conventional vapor compression system  1  of  FIG. 1 , the level of charge is preset and does not vary. As a result, the selected charge must be able to safely operate over the entire range of anticipated sinking temperatures. In an overcharged state the P SD /T SD  can no longer be maintained no matter how much sinking mass flow is provided to the condenser. When this occurs, the system efficiency falls off, as the compressor power consumption increases due to the higher than required compressor discharge pressure. Undercharging a system results in higher energy consumption than optimum due to loss of subcooling and ultimately results in the metering device  30 , such as an expansion valves, being fully open and the eventual loss of control of T L . 
     Total refrigerant mass can be a critical parameter to vapor compression system operational readiness. For example, too much or too little refrigerant can be detrimental to system performance. Extreme values of refrigerant charge can lead to a loss of evaporator temperature control, loss of high-side pressure control, or other potentially catastrophic occurrences. Methods according to embodiments herein, therefore, may relate to real-time methods for determination of acceptable refrigerant charge in a vapor compression system as a function of operational points using only sensors utilized in the control system (i.e., in-situ control sensors). In some embodiments, the methods may incorporate a system that includes a simple in-situ prognostic tool for monitoring a state-of-health (e.g., “Red Light, Yellow Light, Green Light”), with respect to level of charge. Thereby, the system may be maintained on demand, based on the indication from the prognostic tool. Additionally, embodiments herein may relate to methods for continuous management of refrigerant charge as a means for optimizing system efficiency over a range of dynamic operating points. 
     Having described above general configurations and shortcomings of conventional vapor compression systems, vapor compression systems and methods for controlling the systems will now be described according to illustrative, non-limiting embodiments. The systems and methods according to embodiments described herein may mitigate or overcome the disadvantages of the conventional vapor compression system described above. 
     According to some embodiments of methods for controlling vapor compression systems, a system such as a single-load vapor compression system  200  shown in  FIG. 4  may be used. The single-load vapor compression system  200  of  FIG. 4  is only slightly modified from the conventional vapor compression system  1  of  FIG. 1 , but the slight modifications introduce possibilities for cycle optimization that are not inherent in the conventional vapor compression system  1 . 
     The single-load vapor compression system  200  according to the embodiment of  FIG. 4  includes a refrigeration loop  10  configured to transfer heat from a load location  250  to a rejection location  270 . The load location  250  is at a load temperature that results from a heat load, an amount of heat present in a certain volume, at the load location  250 . This heat load defines an evaporator load of the single-load vapor compression system  200 . The refrigeration loop  10  includes several components in fluidic communication with each other through refrigeration lines containing a refrigerant. For example, a compressor  220  is provided that compresses refrigerant vapor from a low-pressure side  12  of the refrigeration loop  10  and delivers compressed refrigerant vapor to a high-pressure side  14  of the refrigeration loop  10 . In various embodiments, the compressor  220  is a fixed capacity compressor. In further embodiments, the compressor  220  is a variable capacity compressor. 
     In the single-load vapor compression system  200 , a condenser  260  is provided at the rejection location  270  and is configured to remove heat from the compressed refrigerant vapor from the compressor  220  to form condensed liquid refrigerant on the high-pressure side  14 . The condenser  260  condenses at least a portion of the refrigerant from the compressor  220  to produce chilled refrigerant, which may include some vapor refrigerant mixed with the condensed refrigerant. In preferred operation the refrigerant is completely condensed exiting the condenser  260  with no vapor refrigerant remaining. The condenser  260  may be in thermal communication with the rejection location  270  via a rejection apparatus  265 . For example, the rejection apparatus may be a fan, a vent, an open cooling loop, a closed cooling loop containing a fluid cooling medium, or a variable bypass path. With such types of rejection apparatus  265 , the rejection capacity may be adjusted, for example, by modifying the fan speed, changing an opening size of the vent, modifying a circulation speed of fluid cooling medium in the closed cooling loop, or modifying the refrigerant flow bypassing the condenser. The rejection apparatus  265  removes sensible and latent heat from the compressed refrigerant vapor and sends the heat to the rejection location  270 . When the rejection apparatus  265  includes a closed cooling loop, the fluid coolant medium may be any thermal-transfer medium such as a liquid (water or glycol, for example) or a gas (air, for example). A closed cooling loop may also include a circulating apparatus such as a pump or other suitable machine for circulating a liquid or vapor condenser fluid. 
     The single-load vapor compression system  200  may further include a charge control loop  210 , through which refrigerant may be added or removed from the refrigeration loop  10 . The charge control loop  210  may include a discharge control valve  216 , a mass storage vessel  212 , and a transfer pump  214 . When control apparatus  290  of the single-load vapor compression system  200  recognizes that the vapor compression system  200  is overcharged and has too much refrigerant in the refrigeration loop  10 , the discharge control valve  216  of the charge control loop  210  may be opened, left open, or opened further to bleed off some of the refrigerant mass. The refrigerant bled off in this manner is then stored in the mass storage vessel  212 . When it is recognized subsequently that the vapor compression system  200  is undercharged and has too little refrigerant in the refrigeration loop  10 , the discharge control valve  216  may be closed, left closed, or closed further, and the transfer pump  214  may be used to inject refrigerant from the mass storage vessel  212  back into the refrigeration loop  10 . Thus, during operation, the charge control loop  210  may be used to actively increase or decrease the mass of refrigerant in the refrigeration loop  10  based on feedback of other sensors in the system that indicate the refrigerant mass is not optimal (either too high or too low). In some embodiments, the charge control loop  210  may include one or more check valves  218 . 
     In the single-load vapor compression system  200 , a metering device  230  is provided, in which condensed liquid refrigerant from the high-pressure side  14  expands and is delivered back to the low-pressure side  12  as chilled refrigerant. In some embodiments, the metering device  230  may be an electronic expansion valve. In other embodiments, the metering device  230  may be a mechanical expansion valve. The condensed liquid refrigerant flows through an adjustable opening of the metering device  230  having an opening width that can be modified as necessary to maintain desired system parameters. Expansion valves, particularly the adjustable openings of expansion valves, may be controlled by a suitable control apparatus to introduce a calculated or predetermined amount of refrigerant into the evaporator  240  by controlling how widely the metering device  230  is open. The amount of refrigerant introduced into the evaporator  240  may be adjusted using the metering device  230  during operation of the single-load vapor compression system. 
     In the single-load vapor compression system  200 , an evaporator  240  is provided at the load location  250  and is configured to transfer heat from the load location  250  to the chilled refrigerant from the metering device  230  to form the refrigerant vapor on the low-pressure side  12  for recirculation into the compressor  220 . This transfer of heat warms the refrigerant while cooling the space surrounding the load location  250 . Circulating means such as a fan may be present at the load location  250  to enhance cooling of the space surrounding the load location  250 . 
     The single-load vapor compression system  200  may include sensors or control apparatus. For example, a low-side sensor  237  may be disposed between the evaporator  240  and the compressor  220  for measuring a low-side pressure or a suction condition (such as P S , T S , or both) of the low-pressure side  12  of the refrigeration loop  10 . A high-side sensor  280  may be disposed between the condenser  260  and the metering device  230  for measuring high-side pressure or a condenser discharge condition (such as P D , T CD , or both) of the high-pressure side  14  of the refrigeration loop  10 . A sink sensor  268  may be provided to measure the temperature of coolant entering the condenser  260  or sink temperature (T CI ). A load sensor  242  may be provided to measure an output temperature or load temperature T L  at the load location of the evaporator  240 . A control apparatus  290  may be electronically coupled to the compressor  220 , the metering device  230 , the low-side sensor  237 , the high-side sensor  280 , the load sensor  242 , the rejection apparatus  265 , the transfer pump  214 , and the discharge control valve  216  or a subset thereof. The control apparatus  290  may include a computer or processor that is capable of sending electronic signals to each of the connected components, whereby the electronic signals instruct the connected components to perform mechanically or otherwise according to specifications of the manufacturers of the components. It should be understood that the positions of sensors in  FIG. 4  are intended to illustrate one option for the positions and that the sensors may be moved to other suitable locations at which the same pressure and temperature values can be determined. 
     In some embodiments, the single-load vapor compression system  200  may include a liquid injection valve  282  that cycles compressed liquid refrigerant back to the refrigerant vapor on the low-pressure side  12 . When present, the liquid injection valve  282  expands the compressed liquid refrigerant, which is immediately chilled, to moderate the temperature of any superheated refrigerant vapor before the superheated refrigerant vapor enters the compressor  220 . Moderation of the temperature of the superheated refrigerant vapor may be used to ensure the compressor  220  operates with an inlet temperature within the manufacturer&#39;s specifications. 
     In some embodiments of methods for controlling the single-load vapor compression system  200 , the methods may include selecting a desired set-point temperature range for the evaporator  240 , as measurable by the load sensor  242 . The desired set-point temperature may be selected manually or may be entered into the control apparatus  290 . As illustrative examples, the set-point temperature range may be selected as 30° C.±5° C., or 10° C.±3° C., or −50° C.±0.2° C. Once the desired set-point temperature range is selected, the methods may further include continually adjusting with the control apparatus  290 : a capacity of the compressor  220 , the adjustable rejection capacity of the rejection apparatus  265 , the adjustable opening of the metering device  230 , and the refrigerant mass in the refrigeration loop  10  with the charge control loop  210  via the transfer pump  214  and discharge control valve  216 . 
     According to some embodiments, the capacity of the compressor  220  may be adjusted to maintain a maximum low-side pressure on the low-pressure side  12  of the refrigeration loop  10  as measured by the low-side sensor  237  while also maintaining the desired set-point temperature at the evaporator  240 . In some embodiments, the maximum low-side pressure is also the maximum saturated-suction pressure. The maximum low-side pressure may be maintained, for example, by adjusting compressor speed to attain a minimum temperature difference threshold (such as5° C. or 10° C., for example) between the desired set-point temperature at the load sensor  242  and the temperature of refrigerant entering the inlet of the compressor  220 , i.e., the suction temperature T S , as determined by the low-side sensor  237 . The maximum low-side pressure possible on the low-pressure side  12  is dependent in part on the evaporator load set point temperature; because for heat to transfer from the load location  250  to the refrigeration loop  10  in the evaporator  240 , the temperature of the refrigerant in the evaporator  240  must be lower than the temperature of the load location. A higher saturated suction pressure P SS , for example, translates to a higher suction temperature T S . 
     According to some embodiments, the rejection apparatus  265  has an adjustable capacity which may be adjusted so as to maintain a minimum high-side pressure as measured by the high-side sensor  280 . In some embodiments, the minimum high-side pressure is also the minimum saturated-discharge pressure. The high-side pressure is affected by the amount of heat removed from the refrigerant via the rejection apparatus  265 . Therefore, the high-side pressure may be decreased generally by increasing a fan speed, opening a vent more widely, or increasing circulation of a fluid cooling medium through a closed cooling loop, for example, depending on the type of rejection apparatus  265  present. In general, with closed cooling loops a higher circulation speed of a circulating apparatus such as a pump may result in a lower high-side pressure, and a lower circulation speed of the circulating apparatus may result in a higher high-side pressure. 
     In the conventional vapor compression system  1  the minimum high-side pressure attainable may be constrained by a maximum allowable opening of the metering device  30 . However, the charge control loop  210  allows for refrigerant to be added to the refrigeration loop  10  resulting in the ability for the metering device  230  to close down. 
     According to some embodiments, the adjustable opening of the metering device  230  may be adjusted so as to maintain the output temperature or load temperature T L  (measured by the load sensor  242 ) within the desired set-point temperature range. If the load sensor  242  measures load temperature T L  above the desired set-point temperature range, for example, the metering device  230 , such as the adjustable opening of an expansion valve, may be opened more widely to increase the refrigeration flow and, thereby, decrease the load temperature T L . If the load sensor  242  measures a load temperature T L  below the desired set-point temperature range, for example, the metering device  230  may be closed slightly to decrease the refrigerant flow and, thereby, decrease the load temperature T L . It may be preferable that the pressure drop across the metering device  230  be held relatively constant in contrast to the percentage opening of the metering device  230 . In some embodiments, the single-load vapor compression system  200  may be configured such that the adjustable opening of the metering device  230  remains continually open to its widest extent, such that attainment of the desired set-point temperature range may occur by adjusting only the capacity of the compressor  220  to modulate P S  and the adjustable rejection capacity of the rejection apparatus  265  to minimize P D . In such embodiments, adjusting the adjustable opening of the metering device  230  may comprise only a single adjustment of the adjustable opening of the metering device  230  to its full-open position. 
     According to some embodiments, the charge or amount of refrigerant in the refrigeration loop  10  may be adjusted to maintain a minimum P SD /T CD  as measured by the high-side sensor  280  based on the available sink temperature. During operation of the vapor compression system  200  of  FIG. 4 , when the system recognizes that the P D  is too high with respect to known or desired optimal operating conditions, the discharge control valve  216  of the charge control loop  210  may be opened, left open, or opened further to bleed off some of the refrigerant volume from the refrigeration loop  10  into the mass storage vessel  212 . When it is recognized subsequently that the T CD  is too high with respect to known or desired optimal operating conditions, the discharge control valve  216  may be closed, left closed, or closed further, and the transfer pump  214  may be used to inject refrigerant from the mass storage vessel  212  back into the refrigeration loop  10  at an appropriate pressure. Thus, during operation, the charge control loop  210  may be used to actively increase or decrease the mass of refrigerant in the refrigeration loop  10 , based on feedback of other sensors, such as the high-side sensor  280 , that indicate the refrigerant mass is either too high or too low. 
     In an exemplary method for operating or controlling the vapor compression system  200  of  FIG. 4 , during normal operation of the system, refrigerant mass may be increased until the T CD  is within about5° F. of the inlet coolant or sinking temperature. The difference between T CD  and the temperature of the coolant entering the condenser at the sink temperature (T CI ) is referred to as the approach for purposes of this disclosure and an approach high set-point is a maximum target approach. Non-limiting further examples of an approach high set-point include temperature differentials between T CD  and T CI  of 4° F., 7° F., 9° F., 12° F., and 15° F. Alternatively, the refrigerant mass may be increased if a higher T SD  is desired due to other constraints such having any metering device  230  in the vapor compression system  200  exceeding the maximum steady-state recommended percent open. Non-limiting examples of the maximum steady-state recommended percent open for the metering device  230  include 90%, 85%, 80%, 75%, 70%, 65%, 60%, and 55%. During the operation, the active control of refrigerant mass through the charge control loop  210  may further include decreasing the refrigerant mass, in the manner described above, if the approach is below an approach low set-point, representing a minimum target approach, of less than 1° F., less than 2° F., less than 3° F., less than 4° F., or less than5° F., for example. Additionally, the active control of refrigerant mass through the charge control loop  210  may further include decreasing the refrigerant mass, in the manner described above, if a lower T SD /P SD  is desired such as if the T SD  exceeds a saturated discharge temperature set-point. Non-limiting examples of the saturated discharge temperature set-point include saturated discharge temperatures which exceed a desired saturated discharge temperature by 1° F., 2° F., 3° F., 4° F., and5° F. In various embodiments, decreasing the refrigerant mass is only allowable if the most open expansion valve is sufficiently closed such as less than 65% open, less than 70% open, less than 75% open, or less than 80% open. 
     In an exemplary method for operating or controlling the vapor compression system  200  of  FIG. 4 , determination by the control apparatus  290  of the desirability of adding or removing refrigerant from the refrigeration loop  10  is made on a periodic basis at polling intervals. In various embodiments the control apparatus  290  polls the sensors approximately every thirty seconds and adds refrigerant, removes refrigerant, or maintains refrigerant charge in the refrigeration loop  10  steady based on the sensor readings. Non-limiting examples of control apparatus  290  polling frequency in further embodiments include approximately 15 second intervals, approximately 1-minute intervals, approximately 30-minute intervals, approximately 1-hour intervals, approximately 12-hour intervals, and approximately 7-day intervals. For example, polling at 15-second intervals may be desirable when T CI  is changing rapidly or 7-day intervals may be desirable when infrequent monitoring of system health is warranted. Additionally, in some embodiments the polling interval of the control apparatus  290  may be changed depending on the operational environment of the vapor compression system. For example, an aircraft cooling system may poll the control apparatus  290  every 15 seconds during ascent and decent and only once every 1 minute during cruising operations at steady altitude. Similarly, a home HVAC system may poll the control apparatus  290  every 5 minutes during daytime hours but only once an hour during evening or nighttime hours. 
     In an exemplary method for operating or controlling the vapor compression system  200  of  FIG. 4 , the control apparatus  290  may add from about 0.25 pounds of refrigerant to about 10 pounds of refrigerant or remove about 0.25 pounds of refrigerant to about 10 pounds of refrigerant each time the results of the polling of the sensors dictate a charge adjustment. In further embodiments, less than 0.25 pounds of refrigerant, about 0.25 pounds of refrigerant, about 0.5 pounds of refrigerant, about 1 pound of refrigerant, about 2 pounds of refrigerant, about 5 pounds of refrigerant, about 10 pounds of refrigerant, or more than 10 pounds of refrigerant may be added or removed each time the results of the polling of the sensors dictate a charge adjustment. One having skill in the art should recognize that the amount of refrigerant added or removed is dependent on the size and configuration of the vapor compression system and should appreciate that a vast array of adjustments are plausible. Additionally, the amount of refrigerant added or removed during each polling cycle by the control apparatus  290  may vary depending on the magnitude by which the control apparatus  290  determines the present level of refrigerant in the refrigeration loop  10  needs to be adjusted. Additionally, refrigerant added or removed may be determined by a PD loop (proportional derivative controller) in various embodiments. 
     Without an active charge control system as a result of the charge control loop  210 , the vapor compression system  200  would have to be charged to a level that provided safe operation over the whole range of environmental conditions, thus potentially being a compromise or less than optimal for at least some conditions, or even a majority of conditions, encountered during normal operations. In the field of aircraft operation, for example, the selection of the median fixed charge must be suitable for operation on the ground during a hot day and at a substantially colder high altitude. In such a scenario, the safe charge would favor the ground hot day operation and would penalize the cold high altitude. If the system were charged for the cold high-altitude operation, there would be excessive charge for the hot-day ground operation. Additionally, aircraft spend a higher percentage of operating time at a high altitude cold operation than on the ground, so this penalty for a fixed charge would accumulate over time in the total amount of engine fuel burned and ultimately would limit the aircraft range and increase the cost of operation. The ability to optimize the charge for both operation on the ground during a hot day and cold high-altitude operation by adjusting the charge in the refrigeration loop  10 , allows the vapor compression system  200  to have an overall higher operating efficiency. 
     Additionally, in lieu of automatic adjustment of the charge of refrigerant in the refrigeration loop  10 , the control apparatus  290  may provide notification in the form of a prognostic tool of the state-of-health of the system. If the control apparatus  290  determines the level of charge is too high or too low the control apparatus  290  may alert the condition with a light or similar indicator to notify an operator to manually correct the refrigerant level. The light may be, for example, a green light to indicate normal charge level, a yellow light to indicate marginal charge level, and a red light to indicate problematic charge level. Alternatively a series of lights or other indicators may indicate low charge, marginal charge, adequate charge, and high charge through a series of 4 or more indicators. 
     In some embodiments, the optimization strategies of the above-described methods for controlling single-load vapor compression systems may apply directly to methods for controlling multiple-load vapor compression systems. A schematic diagram of an exemplary embodiment of a multiple-load vapor compression system  300  is provided in  FIG. 5 . As in the single-load vapor compression systems described above, the multiple-load vapor compression system  300  may include a compressor  220  and a condenser  260 . Though not shown in  FIG. 5 , multiple-load vapor compression systems having more than one compressor or more than one condenser are also contemplated. The multiple-load vapor compression system  300  includes more than one evaporator. Though for sake of simplicity the schematic of  FIG. 5  includes only two evaporators, a first evaporator  240 A and a second evaporator  240 B connected in parallel to the refrigeration loop, additional evaporators may be present in any desired parallel configuration. In some embodiments, the multiple-load vapor compression system  300  may include more than two evaporators, for example up to five evaporators, up to ten evaporators, or more than ten evaporators. The multiple-load vapor compression system  300  also includes an independently controllable metering device, such as an expansion valve, in line with each evaporator. For example, a first metering device  230 A is in line with the first evaporator  240 A, and a second metering device  230 B is in line with the second evaporator  240 B. 
     Analogously to the single-load vapor compression systems, the multiple-load vapor compression system  300  moves heat from both a first load location  250 A and a second load location  250 B to a rejection location  270 . The first load location  250 A and the second load location  250 B may have the same ambient temperatures, nearly the same ambient temperatures (such as ±5° C.) compared to each other, or very different ambient temperatures (such as ±20° C., ±50° C., or even ±100° C.) compared to each other. The first load location  250 A and the second load location  250 B may or may not have any thermal communication with each other outside the multiple-load vapor compression system  300  itself. For example, the first load location  250 A may be a habitable room or a refrigerator in a building and the second load location  250 B may be a deep freezer in the same building. When the temperatures at the first load location  250 A and the second load location  250 B are different, they may be said to represent different independent thermal loads on the multiple-load vapor compression system  300 . 
     The first load location  250 A is in thermal communication with the first evaporator  240 A through the first evaporator loop  245 A, and the second load location  250 B is in thermal communication with the second evaporator  240 B through the second evaporator loop  245 B. The first evaporator has a first-evaporator output temperature T L1  that may be measured by first-evaporator sensor  242 A, and the second evaporator has a second-evaporator output temperature T L2  that may be measured by second-evaporator sensor  242 B. 
     The control apparatus  290  of the multiple-load vapor compression system  300  may be electronically coupled to the compressor  220 , the rejection apparatus  265 , the transfer pump  214 , the discharge control valve  216 , the first-evaporator sensor  242 A, the first metering device  230 A, the second-evaporator sensor  242 B, the second metering device  230 B, and the liquid injection valve  282  (when present). 
     In exemplary embodiments of methods for optimized control of the multiple-load vapor compression system  300 , a first desired set point output temperature T L1  for the first evaporator  240 A may be selected, and a second desired set point output temperature T L2  for the second evaporator  240 B may be selected. When more than two evaporators are present, set points for each additional evaporator may also be selected. 
     As with the single-load vapor compression system, the control methods may further include minimizing the saturated discharge condition (such as P SD , T SD , or both) while maximizing the saturated suction condition (such as P SD , T SD , or both) within the operating constraints of the multiple-load vapor compression system  300 , particularly within the operating constraints of the compressor  220 . The control apparatus  290  may then check the saturated suction condition and the saturated discharge condition against the output temperatures of the evaporators and independently adjust the output temperatures by controlling the state of the expansion valve associated with each evaporator. For example, if the measured output temperature T L1  of the first evaporator is too high, the control apparatus  290  may direct the first metering device  230 A to increase the flow of refrigerant toward the first evaporator  240 A. Additionally, in conjunction with adjustment of the first metering device  230 A and/or the second metering device  230 B, the control apparatus  290  may increase or decrease the charge in the refrigeration loop  10  by adding or removing refrigerant from the refrigeration loop  10  through control of the transfer pump  214  and/or the discharge control valve  216  to ensure all output temperatures (T L1 , T L2 , for example) at all evaporators  240 A,  240 B are within an acceptable range of their desired set-points. Alternatively, the control apparatus  290  may increase or decrease the cycle time and/or the speed of the compressor  220  to ensure all output temperatures (T L1 , T L2 , for example) at all evaporators  240 A,  240 B are within an acceptable range of their desired set-points. In various embodiments, the multiple-load vapor compression system  300  includes back pressure valves (not shown). 
     In some embodiments, the control apparatus  290  may be programmed with one or multiple algorithms that adjust any or all operating parameters in a manner that ensures (a) that the high-side pressure is minimized; while (b) the low-side pressure is maximized; while (c) each output temperature of each evaporator is within the desired set-point output temperature range. 
     With reference to  FIG. 6 , an exemplary active charge control scheme is described in accordance with an embodiment of the present disclosure. Initially the condenser coolant flow is determined and compared to a target flow. If the condenser coolant flow is below the target flow the position of the expansion valves (metering device) is reviewed. If the condenser coolant flow is not below the target flow the T SD  is reviewed. In reviewing the position of the expansion valves, if none of the expansion valves are opened beyond approximately 60% the T SD  is reviewed. Conversely, if in reviewing the position of the expansion valves any expansion valve is opened beyond approximately 60% the superheat at evaporator exit and subcooling at condenser exit are reviewed. If superheat at the evaporator exit exceeds approximately 10° F. or subcooling at the condenser exit falls below approximately 10° F., charge is added to the refrigeration loop. Conversely, if superheat at the evaporator exit does not exceed approximately 10° F., the control system waits a polling interval, such as 30 seconds, and repeats the analysis from the beginning. If subcooling at the condenser exit exceeds approximately 10° F., the T SD  is reviewed. In reviewing the T SD , the T SD  is compared to the saturated discharge temperature set-point. If the T SD  exceeds the saturated discharge temperature set-point, charge is removed from the refrigeration loop. Conversely, if the T SD  remains below the saturated discharge temperature set-point, the temperature differential between the coolant inlet temperature and the refrigerant discharge temperature of the condenser, known as the approach, is reviewed. If the approach exceeds the approach high set-point, charge is added to the refrigeration loop. If the approach remains below the approach high set-point, the approach is compared to the approach low set-point. If the approach remains above the approach low set-point, the control system waits a polling interval and repeats the analysis from the beginning. Conversely, if the approach drops below the approach low set-point, charge is removed from the refrigeration low. Returning once again to the T SD  review, if the T SD  exceeds the saturated discharge temperature set-point, charge is removed from the refrigeration loop. Finally, after charge is added or removed from the refrigeration loop the control system waits a polling interval and repeats the analysis from the beginning. One having ordinary skill in the art would appreciate that the threshold values provided in this explanation of  FIG. 6  are exemplary in nature and alternate values are also applicable. 
     Further embodiments include cooling systems; refrigeration systems; home HVAC systems; automotive or vehicular cabin or cockpit cooling systems; aircraft fuel tank cooling systems; or other similar systems incorporating the components described herein and operated according to the methods described herein. Such systems may include one or multiple heat loads, and when multiple heat loads are present, the multiple heat loads may be dissimilar or substantially the same. It is also contemplated that such systems may include one or multiple compressors, and that the one or more compressors may include multi-stage or inner-stage compressors. 
     Further embodiments herein include any of the vapor compression systems described above, for which the load location or at least one of multiple load locations is an enclosed space. When multiple load locations are present, each load location may be an enclosed space that is cooled independently by the expansion valves of the vapor compression system. In illustrative embodiments, load locations may include any enclosed space in radar apparatus, an aircraft, an electronic apparatus, a cabin environment, a cockpit environment, a weapon, a galley, a fluidic apparatus containing lubrication fluids, and a fuel compartment containing a fuel, for example. 
     Further embodiments herein include any of the vapor compression systems described above, for which the rejection location is the environment. In other embodiments, the rejection location of a particular vapor compression system may be an intermediate rejection location from which additional heat is removable to an ultimate rejection location. In such embodiments, as illustrative non-limiting examples the intermediate rejection location may include chilled water, a fuel tank, an air stream, or a body of water. The ultimate rejection location may be the environment. 
     EXAMPLES 
     To determine the effect of modifying the charge of refrigerant in the refrigeration loop  10  experiments were run with the independent variables including refrigerant mass charge, heat load, T SD  set point, condenser cooling fluid inlet temperature, and the T L  set point. The experiments were run on an exemplary vapor compression system and are not meant to be limiting to the variety of vapor compression systems covered by the present disclosure. Exemplary components include a compressor available from Fairchild Corporation, a Danfoss 70 kW condenser (B3-095-72-H), Emerson expansion valves, and a Yaskawa A-1000 controller. These tests allowed the impact of refrigerant charge on vapor compression system efficiency measured as coefficient of performance (COP) and ability to maintain T L  within a desired margin, for example 2° F., of the set point to be accessed. To accomplish this, the refrigerant charge was systematically changed and system performance was characterized. 
     An experimental set-up of the vapor compression system with an adjustable heat load was allowed to stabilize at a constant evaporator load of 24 kW, with the T L  set point fixed at 65° F. Refrigerant charge was then slowly added or removed from the system in 1-pound or 2-pound increments and was allowed to stabilize, at which point system performance was characterized. This was repeated at three T SD  settings and six different condenser inlet coolant temperatures. In this manner it was possible to examine the relationships of charge to COP, compressor speed, high side pressure, condenser refrigerant exit temperature, subcooling, expansion-valve position, and the evaporator-outlet load temperature. 
     It is generally accepted that increasing subcooling increases the COP to a limit. The limit is generally accepted as an unwanted increase in high side pressure. This increase in COP is the result of an increase of available enthalpy change across the evaporator(s), which decreases the required mass flow and thus reduces the compressor speed needed to maintain the evaporator-load temperature(s). 
     The measured relationship between condenser discharge temperature and COP was determined with the experimental set-up of the vapor compression system. The experimental set-up of the vapor compression system was operated with a constant load of 24 kW, a T SD  set point of 135° F., and a condenser inlet sink temperature of 80° F. The condenser discharge temperature decreased to the limit of the sink temperature as the charge was increased from 22 pounds to 32 pounds. Throughout this range of charge the subcooling likewise increased. Addition of refrigerant beyond 32 pounds of charge resulted in negligible change in condenser discharge temperature or subcooling. The COP increased from 22 pounds to 32 pounds of charge and then remained relatively constant from 32 pounds to 36 pounds. However, above 36 pounds of charge the COP begins to decrease and the subcooling began to increase. This COP decrease and subcooling increase was attributed an increase in high side pressure and inability to maintain T SD  or high side pressure. 
     The COP decrease and subcooling increase above 36 pounds of charge was attributed an increase in high side pressure and inability to maintain T SD  or high side pressure. This increase in high side pressure corresponds with an increase in condenser coolant flow. Additionally, above 37 pounds of charge the coolant flow was integrating to maximum flow from the T SD  control loop. This is a direct result of the control loop not being able to contain the high-side pressure. The condenser heat transfer is limited by one of three factors: the heat capacity of the coolant ({dot over (m)}C p ) or phase change enthalpy of the refrigerant ({dot over (m)}H fg ) or product of the condensing area (A) and overall heat transfer coefficient (U) of the heat exchanger itself. The condensing capacity was sensitive to coolant flow from 22 pounds to 36 pounds of charge. In this region the limiting factor in the condenser heat transfer was the heat capacity of the coolant ({dot over (m)}C p ). Above 36 pounds, the limiting factor appears to be the condensing area (A). This insufficient condensing area results in the loss of T SD  control, which appears as both an increase in T SD  and high-side pressure. The increase in high-side pressure results in an increase in compressor pressure rise and, thus an increase in work, resulting in a reduction in COP. 
     Conversely, there are detrimental effects from having too little charge. Below about 23 pounds of charge the experimental set-up of the vapor compression system lost the ability to control T L . Once the charge dropped below 25 pounds the compressor speed became more erratic and even though the experimental set-up of the vapor compression system was able to maintain the load temperature control at 25 pounds of charge, it was becoming marginal. It is believed the experimental set-up of the vapor compression system had insufficient charge when the expansion valve position was greater than about 50% but this value is dependent on the expansion valve size selected and the load. In the experimental set-up of the vapor compression system the expansion valve was oversized for the load. This parameter can provide clear indication of too little charge and incipient loss of control. 
     By combining the various tests, one can conclude that the experimental set-up of the vapor compression system will stay within regulation between 23 pounds and 36 pounds of charge. In practice, it is impractical to know the absolute charge without physically evacuating it and recharging the system. Additional the actual charge values will vary depending on individual system configuration, size, and other physical parameters. It is therefore desirable to discuss charge management in terms of either addition or subtraction of charge. The leading signs of low charge are high percentage opening of the expansion valve, and loss of subcooling. Low subcooling can also be attributed to low temperature difference between T SD  and the condenser sink temperature rather than low charge. 
     The experimental set-up of the vapor compression system was run and the effects of system charge and sink temperature at a constant T SD  was determined. Measured COP values were determined as a function of charge at a constant T SD  set point of 135° F. but at 3 condenser sink temperatures of 110° F., 65° F., and 20° F. As the charge increased, T SD  control is lost and there is a very significant drop in COP. It appears that the drop in COP becomes more acute as the charge increases, which supports the loss of condensing area with increasing charge. 
     Additionally the allowable charge range varied with the differences between T SD  and the sink temperatures. At small temperature differences, such as the 110° F. sink temperature (25° F. temperature difference), there is only a 3-pound window of charge in which the system could be safely operated, whereas at a 20° F. sink temperature (115° F. temperature difference) the charge range is nearly 24 pounds. If the experimental set-up of the vapor compression system had to operate with a fixed charge over this range of sink temperatures with this T SD  it would need to be from 16 to 18 pounds to accommodate the 110° F. sink temperature. 
     As is evident from the discussion above, operation of the experimental set-up of the vapor compression system at maximum efficiency may include running the system at the lowest T SD  possible based on the available sink temperature. As acceptable operating charge is also a function of the T SD  and sink temperature adjustability of the charge in the refrigeration loop is desirable. This is best observed at the high sink temperature of 110° F., where there is a narrow range of only a few pounds of charge between loss of load control and loss of T SD  or high-side pressure control. As this temperature difference increases, the allowable charge range also increases. At a 65° F. sink temperature the allowable charge range increases to 13 pounds. The sink temperature of airborne applications can vary widely from −60° F. to over 100° F. thus making it very advantageous to dynamically adjust the system charge to achieve peak COP at minimum T SD . 
     As the experiments detailed above show, charge affects nearly every aspect of the operation of a vapor compression system. From the data acquired, three distinct operating regions can be delineated: too high of charge, too low of charge, and an acceptable range that may include a marginal charge or an adequate charge. By splitting the acceptable range into two states, marginal or adequate, four states may be defined: low charge, marginal charge, adequate charge, and high charge. 
     The low charge state may be characterized by a loss of load control, excessive expansion valve opening, and lack of subcooling, for example. The detection of incipient failure is loss of subcooling, and excessive expansion valve opening exceeding about 50%, which varies depending on the evaporator load and relative expansion valve capacity. The marginal charge state may be characterized by expansion valve openings greater than nominal but less than the 50% limit, and small difference between the sink temperature and the condenser exit temperature. The adequate charge state may be characterized by maintained T SD  control and load temperature control. The high charge state may be characterized by inability to maintain T SD  control, resulting in excessive high-side pressures. The subcooling can exceed the desired level when the actual T SD  exceeds the T SD  set point, while the refrigerant discharge temperature remains constant. The data also suggest it may be advantageous in terms of both operability and efficiency to have an active charge control system for vapor control systems system, in which wide ranges of sink temperatures are present. Nevertheless, it should be understood that the active charge control configurations and operation methods apply to all vapor compression systems, such as household or industrial HVAC systems, refrigeration systems, or heat pumps, for example. 
     Unless otherwise defined, all technical and scientific terms used herein have the same meaning as commonly understood by one of ordinary skill in the art to which the claimed subject matter belongs. The terminology used in the description herein is for describing particular embodiments only and is not intended to be limiting. As used in the specification and appended claims, the singular forms “a,” “an,” and “the” are intended to include the plural forms as well, unless the context clearly indicates otherwise. 
     It is noted that terms like “preferably,” “commonly,” and “typically” are not utilized herein to limit the scope of the appended claims or to imply that certain features are critical, essential, or even important to the structure or function of the claimed subject matter. Rather, these terms are merely intended to highlight alternative or additional features that may or may not be utilized in a particular embodiment.