ENERGY-OPTIMIZED CYCLE CONTROL OF TIME-VARIANT LOADS FOR THERMAL MANAGEMENT OF VAPOR COMPRESSION SYSTEMS

[0001] This application claims the benefit of priority under 35 U.S.C. §119(e) to U.S. Provisional Application Ser. No. 61/751,554, filed Jan. 11, 2013.

[0002] This invention was made with government support under contract FA8650-04-D-2403-0017, awarded by The Air Force Research Laboratories (AFRL). The government has certain rights in the invention.

[0003] The present disclosure relates generally to vapor compression systems and, more particularly, to methods for controlling the cycle of vapor compression systems to optimize energy usage and thermal management.

[0004] 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.

[0005] Conventional methods for thermal management and process control of vapor compression systems are based on decades-old practices that were established when technology, components, and computerized controllers were either non-existent or far less sophisticated than they are presently. Though the conventional processes may accomplish the thermal management tasks they set out to accomplish, they needlessly waste energy because their designs were built around old equipment capabilities. Moreover, passive mechanical valves such as thermal expansion valves are commonly used in multi-load systems, in which a single cooling network may be in communication with multiple evaporators that each may be located in different cooling spaces have different cooling requirements. As such, there are ongoing needs for energy-efficient methods for controlling vapor compression systems, small and large.

[0006] 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 location. The load location has a load temperature resulting from a heat load at the load location. The heat load defines an evaporator load. The refrigeration loop includes a plurality of components in fluidic communication through refrigeration lines containing a refrigerant. The plurality of components include an adjustable-capacity compressor that compresses the refrigerant from a low-pressure side of the refrigeration loop and delivers the refrigerant to a high-pressure side of the refrigeration loop. The plurality of components also includes a condenser that condenses at least a portion the refrigerant from the adjustable-capacity compressor to produce chilled refrigerant. The condenser is in thermal communication with the rejection location via a rejection apparatus having an adjustable rejection capacity. The plurality of components also includes an expansion valve having an adjustable opening through which the chilled refrigerant from the condenser expands and is delivered back to the low-pressure side. The plurality of components also includes an evaporator at the load location that transfers heat from the heat load to the refrigerant arriving from the expansion valve and delivers the refrigerant back to the adjustable-capacity compressor. In addition to the refrigeration loop, the vapor compression system also includes a first sensor that measures a low-side pressure of the low-pressure side of the refrigeration loop; a second sensor that measures a high-side pressure of the high-pressure side of the refrigeration loop; and a third sensor that measures the load temperature. The vapor compression system includes a control apparatus electronically coupled to the adjustable-capacity compressor, the expansion valve, the first sensor, the second sensor, the third sensor, and the rejection apparatus. Thus, the methods for controlling the vapor compression system may include selecting a desired set-point temperature range for the load temperature and then continually adjusting at least three parameters with the control apparatus. One parameter to adjust may be a capacity of the adjustable-capacity compressor so as to maintain with respect to the evaporator load a maximum low-side pressure as measured by the first sensor. Another parameter to adjust may be the adjustable rejection capacity of the rejection apparatus so as to maintain a minimum high-side pressure as measured by the second sensor. Yet another parameter to adjust may be the adjustable opening of the expansion valve so as to maintain the load temperature measured by the third sensor within the desired set-point temperature range.

[0007] According to other embodiments, methods for controlling a multiple-load vapor compression system may include operating a multiple-load vapor compression system. The multiple-load vapor compression system may include a refrigeration loop configured to transfer heat from multiple load locations to at least one rejection location, each load location having a load temperature resulting from a heat load at the load location, the heat load defining an evaporator load for the load location. The refrigeration loop includes a plurality of components in fluidic communication through refrigeration lines containing a refrigerant. The plurality of components include an adjustable-capacity compressor that compresses refrigerant vapor from a low-pressure side of the refrigeration loop and delivers compressed refrigerant vapor to a high-pressure side of the refrigeration loop. The plurality of components also include a condenser that condenses at least a portion of the refrigerant from the adjustable-capacity compressor to produce chilled refrigerant, the condenser being in thermal communication with the rejection location via a rejection apparatus having an adjustable rejection capacity. The plurality of components also include an expansion valve associated with each load location, each expansion valve having an adjustable opening through which the chilled refrigerant from the condenser expands and is delivered back to the low-pressure side. The plurality of components also include an evaporator at each load location that transfers heat from the load location of the evaporator to the refrigerant arriving at the evaporator from the expansion valve associated with the load location and delivers the refrigerant back to the adjustable-capacity compressor. In addition to the refrigeration loop, the multiple-load vapor compression system may also include a first sensor that measures a low-side pressure of the low-pressure side of the refrigeration loop, a second sensor that measures a high-side pressure of the high-pressure side of the refrigeration loop, and multiple third sensors. Each third sensor may be associated with an individual evaporator and may measure the load temperature at the load location of the individual evaporator. The multiple-load vapor compression system may also include a control apparatus coupled to the adjustable-capacity compressor, each expansion valve, the first sensor, the second sensor, each of the third sensors, and the rejection apparatus. Thus, the methods for controlling a multiple-load vapor compression system may include selecting desired set-point temperatures for each load temperature and adjusting at least three parameters continually with the control apparatus. One parameter to be adjusted is a capacity of the adjustable-capacity compressor so as to maintain with respect to the evaporator load at the coldest set point location a maximum low-side pressure as measured by the first sensor. Another parameter to be adjusted is the adjustable rejection capacity of the rejection apparatus so as to maintain a minimum high-side pressure as measured by the second sensor. Yet another parameter to be adjusted is each adjustable opening of each expansion valve independently from other expansion valves in the multi-load vapor compression system, so as to maintain the load temperatures measured by the third sensors associated with each evaporator within the desired set-point temperature range for each load temperature.

[0008] 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.

[0009] 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.

[0016] The methods according for controlling a vapor compression system and for controlling a multiple-load vapor compression system according to various embodiments herein may provide substantial improvements over control methods common to conventional vapor compression systems, particularly with regard to energy efficiency. To provide a contextual basis for the embodiments of the methods, a conventional vapor compression system and its typical control protocol will first be described with reference to FIGS. 1-3. Embodiments of methods for controlling a vapor compression system will be described below with reference to FIGS. 4 and 5. Embodiments of methods for controlling a multiple-load vapor compression system will be described in detail below with reference to FIG. 6.

[0017] A conventional vapor compression system 1 is shown schematically in FIG. 1. 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.

[0018] 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 first sensor 37 is located between the evaporator 40 and the compressor 20. The at least one first 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 saturated suction pressure (P_SS), one or both of which being measurable by the first sensor 37. A saturated suction temperature (T_SS) may be determined from the saturated suction pressure (P_SS) 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's day, the compressed refrigerant vapor must have a temperature greater than 35° C. for the necessary heat exchange to occur.

[0019] Heat is removed from the refrigerant at the condenser 60. The condenser 60 provides communication between the refrigeration lines 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 condenser 60 and the refrigeration lines 10 in the condenser are in thermal communication with a rejection apparatus 65. The rejection apparatus 65 places the condenser 40 in thermal communication with the rejection location 70.

[0020] 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 second sensor 80 between the condenser 60 and the 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 second sensor 80 is called the condenser discharge temperature (T_CD). The pressure of the compressed refrigerant at this stage is saturated discharge pressure (P_SD). A saturated discharge temperature (T_SD) may be determined from the saturated discharge pressure (P_SD) using a pressure-enthalpy table or other suitable table for the refrigerant involved. An amount of subcooling in the system is equal to T_CD−T_SD.

[0021] 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 may 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.

[0022] The chilled refrigerant then travels to the evaporator 40. At the evaporator 40, the refrigeration lines 10 are 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 lines 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 continually being transferred from the load location 50 may be assessed from measuring the output temperature T_L with a third 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.

[0023] 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 10 and an expansion valve 35 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.

[0024] Typically, the expansion valve 35 is designed to adjust the refrigerant flow through the metering device 30 based on the measured temperature difference between the first sensor 37 and the metering device sensor 32. To illustrate, the expansion valve 35 may be set to maintain a threshold difference such as 5° C. Thereby, when the measured temperature difference between the first sensor 37 and the metering device sensor 32 is greater than 5° C., the expansion valve 35 may open more widely to allow more refrigerant to enter the low-pressure side 12. A temperature difference of greater than 5° 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 first sensor 37 and the metering device sensor 32 is less than 5° C., the expansion valve 35 may open less widely to allow less refrigerant to enter the low-pressure side 12. A temperature difference of less than 5° 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.

[0025] 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.

[0026] 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 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.

[0027] 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.

[0028] 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 saturated suction pressure P_SS and is superheated above a saturated suction temperature T_SS (characteristic of mixed-phase refrigerant at pressure P_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 temperature (T_SS) is defined as the temperature of the refrigerant that is inside the dome at the compressor inlet suction pressure (P_SS). In FIG. 3, points 180 and 185 are both are the Pss and Tss. 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 saturated discharge pressure P_SD. 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_SD−P_SS) 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.

[0029] At the condenser, heat is removed from the vapor-phase refrigerant to cause the refrigerant to condense at a constant saturated suction discharge temperature (T_SD) that is characteristic of two-phase refrigerant at pressure P_SD. 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 can be beneficial for prolonging the life of expansion valves and also 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 to expanded liquid 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.

[0030] 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 output temperature T_L measured at the third 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 third 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.

[0031] At the same time a compressor modulation is used in the conventional vapor compression system 1, the conventional vapor compression system 1 seeks to maintain a superheat threshold value between the first sensor 37 measuring saturated suction temperature and the metering device sensor 32 measuring the refrigerant temperature rise across the evaporator. It should be understood that maintenance of the superheat threshold is a maintenance of a temperature difference (T_S−T_MD), not a maintenance of absolute temperatures. Moreover, it should be understood that maintenance of the superheat value is dependent on the pressure drop during adiabatic expansion of compressed liquid refrigerant at the metering device 30. As a result, if the only conventional system constraints are maintenance of output temperature (T_L) and superheat (T_S−T_MD), the constraints may be met by inefficiently running the compressor 20 harder than necessary, thereby compressing vapor refrigerant to a very high pressure and temperature, provided the conventional vapor compression system 1 simply ensures that output temperature (T_L) and superheat (T_S−T_MD) are within acceptable ranges. In other words, even if the compressor 20 wastes energy by compressing the refrigerant to a high pressure and temperature, the conventional vapor compression system 1 continues to operate normally with no indication of system overload or failure. Referring to FIG. 2, the running of the conventional vapor compression system 1 of FIG. 1 results in a pressure increase 190 that is directly related to energy required to run the system and is independent of the superheat margin 187. The waste of energy occurs naturally in the conventional vapor compression system 1, because the saturated discharge temperature (T_SD) and pressure (P_SD) are allowed to float up or down on their own, as long as output temperature (T_L) and superheat (T_SS−T_MD) are within acceptable ranges.

[0032] In addition to the above-noted disadvantages of the conventional vapor compression system 1, it is known that the additional energy that may be required to maintain a superheat margin 187 is never a dominant performance factor. Fairly large swings in superheat margin can have little no effect or negligible effect on heat transfer within the system.

[0033] 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 methods according to embodiments described herein may mitigate or overcome the disadvantages of the conventional vapor compression system described above.

[0034] 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.

[0035] 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, an adjustable-capacity 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. The capacity of the adjustable-capacity compressor 220 may be adjusted by changing the amount of flow produced by the adjustable-capacity compressor 220, by changing the duty cycle of the adjustable-capacity compressor 220 (such as by switching the adjustable-capacity compressor 220 off and on at defined intervals), or both. It should be understood that the term “adjustable-capacity compressor,” therefore, also encompasses a fixed-capacity compressor having a capacity that is adjusted during operation of the single-load vapor compression system 200 by varying duty cycle.

[0036] 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 adjustable-capacity compressor 220 to form condensed liquid refrigerant on the high-pressure side 14. The condenser 260 condenses at least a portion the refrigerant from the adjustable-capacity compressor to produce chilled refrigerant, which may include some vapor refrigerant mixed with the condensed refrigerant. The condenser 260 may be in thermal communication with the rejection location 270 via a rejection apparatus 265 having an adjustable rejection capacity. For example, the rejection apparatus may be a fan, a vent, or 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.

[0037] In the single-load vapor compression system 200, an expansion valve 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 expansion valve 230 may be an electronic expansion valve. In other embodiments, a mechanical expansion valves is contemplated. The condensed liquid refrigerant flows through an adjustable opening of the expansion valve 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 expansion valve 230 is open. The amount of refrigerant introduced into the evaporator 240 may be adjusted using the expansion valve 230 during operation of the single-load vapor compression system.

[0038] In the single-load vapor compression system 200, an evaporator 240 is provided at the load location 250 and configured to transfer heat from the load location 250 to the chilled refrigerant from the expansion valve 230 to form the refrigerant vapor on the low-pressure side 12 for recirculation into the adjustable-capacity 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.

[0039] The single-load vapor compression system 200 may include sensors or control apparatus. For example, a first sensor 237 may be disposed between the evaporator 240 and the adjustable-capacity compressor 220 for measuring a low-side pressure or a saturated suction condition (such as P_SS, T_SS, or both) of the low-pressure side 12 of the refrigeration loop 10. A second sensor 280 may be disposed between the condenser 260 and the expansion valve 230 for measuring high-side pressure or a saturated discharge condition (such as P_SD, T_CD, or both) of the high-pressure side 14 of the refrigeration loop 10. A third 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 adjustable-capacity compressor 220, the expansion valve 230, the first sensor 237, the second sensor 280, the third sensor 242, and the rejection apparatus 265. The control apparatus 290 may be 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 values can be determined.

[0040] 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 adjustable-capacity compressor 220. Moderation of the temperature of the superheated refrigerant vapor may be used to ensure the adjustable-capacity compressor 220 operates with an inlet temperature within the manufacturer's specifications.

[0041] In some embodiments of methods for controlling the single-load vapor compression system 200 may include selecting a desired set-point temperature range for the evaporator 240, as measurable by the third 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 adjustable-capacity compressor 220, the adjustable rejection capacity of the rejection apparatus 265, and the adjustable opening of the expansion valve 230.

[0042] According to some embodiments, the capacity of the adjustable-capacity compressor 220 may be adjusted to maintain respect to the evaporator load a maximum low-side pressure on the low-pressure side 12 of the refrigeration loop 10 as measured by the first 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 compression to attain a minimum temperature difference threshold (such as 5° C. or 10° C., for example) between the desired set-point temperature range at the evaporator 240 and the temperature of refrigerant entering the inlet of the adjustable-capacity compressor 220, i.e., the suction temperature T_S. 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 lines 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.

[0043] The maximum low-side pressure attainable in the system may also be constrained by the maximum allowable opening of the expansion valve 230. The maximum allowable opening represents a condition at which the pressure of expanded liquid refrigerant entering the low-pressure side 12 is higher than when the expansion valve 230 is not fully open, because the overall pressure drop of the liquid refrigerant through the expansion valve 230 will be lower in a full-open state than in a less than fully open state when less refrigerant flows into the low-pressure side 12. The maximum low-side pressure attainable by the maximum opening of the expansion valve 230 may be limited by the minimum allowable pressure drop according to mechanical specifications of the adjustable-capacity compressor 220. In some embodiments, the methods may further include opening the expansion valve 230 to its full-open position or to the maximum open position possible that maintains T_L within an acceptable range of the desired set-point temperature at the evaporator 240, then adjusting the cycle of the adjustable-capacity compressor 220 to maintain the desired set-point temperature at the evaporator 240.

[0044] According to some embodiments, the adjustable rejection capacity of the rejection apparatus 265 may be adjusted so as to maintain a minimum high-side pressure as measured by the second 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.

[0045] The minimum high-side pressure attainable by the single-load vapor compression system 200 may be constrained by a maximum allowable opening of the expansion valve 230. The minimum high-side pressure practical for the single-load vapor compression system 200 may be constrained by the combination of the maximum width of the opening and resulting flow area of the expansion valve 230 and the difference between the P_SD and P_SS. Depending on the specifics of the hardware within the single-load vapor compression system 200, it may be possible to operate the system at a small difference between P_SD and P_SS that results in insufficient refrigerant mass flow at a maximum width of the opening of the expansion valve 230.

[0046] Additionally, the minimum high-side pressure practical for the single-load vapor compression system 200 may be constrained by the cost of the cooling, which cost is related to the energy required to run the adjustable-capacity compressor 220 to compress the mass of refrigerant passing through the adjustable-capacity compressor 220 to the saturated discharge pressure P_SD, particularly of the increased mass flow of refrigerant represented by operating the expansion valve 230 in its full open position. In some embodiments, the pressure rise P_SD−P_SS across the adjustable-capacity compressor may be maintained at or near the lowest value permitted according to the manufacturer's specifications of the adjustable-capacity compressor 220.

[0047] According to some embodiments, the adjustable opening of the expansion valve 230 may be adjusted so as to maintain the output temperature or load temperature T_L (measured by the third sensor 242) within the desired set-point temperature range. If the third sensor 242 measures load temperature T_L above the desired set-point temperature range, for example, the adjustable opening of the expansion valve 230 may be opened more widely to increase the refrigeration flow through 230 and, thereby, decrease the load temperature T_L. If the third sensor 242 measures a load temperature T_L below the desired set-point temperature range, for example, the adjustable opening of the expansion valve 230 may be closed slightly to decrease the refrigerant flow through the expansion valve 230 and, thereby, decrease the load temperature T_L. It may be preferable that the pressure drop across the expansion valve 230 be held relatively constant in comparison to the valve opening. In some embodiments, the single-load vapor compression system 200 may be configured such that the adjustable opening of the expansion valve 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 adjustable-capacity compressor 220 to maximize P_SS and the adjustable rejection capacity of the rejection apparatus 265 to minimize P_SD. In such embodiments, adjusting the adjustable opening of the expansion valve 230 may comprise only a single adjustment of the adjustable opening of the electronic valve 230 to its full-open position.

[0048] In some embodiments, the methods for controlling the single-load vapor compression system 200 may further include optimizing the saturated discharge condition and the saturated suction condition based on a value function. The value function may define the relative cost of cooling at the minimum P_SD to the amount of energy required to operate the adjustable-capacity compressor 220 at the speed that attains the minimum P_SD.

[0049] An operational schematic 400 of the single-load vapor compression system 200 controlled according to the methods described above is provided in FIG. 5 and can be compared with the operational schematic 101 in FIG. 3 of the conventional vapor compression system 1. The pressure-enthalpy curve 405 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 460, the refrigerant is a single-phase vapor at saturated suction pressure P_SS and is superheated above a saturated suction temperature T_SS 485 (characteristic of mixed-phase refrigerant at pressure P_SS) by a superheat margin 487 to ensure no liquid refrigerant enters the compressor. Unlike in the conventional vapor compression system, the superheat margin is not a control parameter. The superheat margin 487 need only be greater than zero. The vapor-phase refrigerant is compressed by the compressor to post-compression point 465 at saturated discharge pressure P_SD. The line from the pre-compression point 460 to the post-compression point 465 follows a path dictated by the compressor, for example near a line of constant entropy. According to the methods described above, P_SD is minimized and P_SS is maximized, and, therefore, the pressure increase 490 (P_SD−P_SS) involved with this process is itself minimized. Because the pressure increase 490 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 optimization parameters included in the methods herein may result in decreased energy consumption and cost.

[0050] 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_SD. The refrigerant condenses to saturation point 470 and then may be subcooled below T_SD by a subcooling margin 472 to reach a subcooled liquid point 475. The liquid refrigerant is then allowed to expand adiabatically through the expansion valve to expanded liquid point 477. This expansion occurs at a constant temperature until the liquid begins to vaporize, at which point further expansion results in a decrease of temperature. As depicted in FIG. 5 it is also possible that none of the liquid ever flashes to vapor through the expansion valve, with the fluid pressure of P_SD at 475 being reduced to P_SS at point 477. Then, at the evaporator the liquid refrigerant vaporizes at a constant temperature to saturated vapor point 485. As noted above, additional heat may be added to the saturated vapor through the superheat margin 487 until the refrigeration cycle again reaches the pre-compression point 460.

[0051] 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. 6. As in the single-load vapor compression systems described above, the multiple-load vapor compression system 300 may include a adjustable-capacity compressor 220 and a condenser 260. Though not shown in FIG. 6, 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. 6 includes only two evaporators, a first evaporator 240A and a second evaporator 240B 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 expansion valve in line with each evaporator. For example, a first expansion valve 230A is in line with the first evaporator 240A, and a second expansion valve 230B is in line with the second evaporator 240B.

[0052] Analogously to the single-load vapor compression systems, the multiple-load vapor compression system 300 moves heat from both a first load location 250A and a second load location 250B to a rejection location 270. The first load location 250A and the second load location 250B 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 250A and the second load location 250B 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 250A may be a habitable room or a refrigerator in a building and the second load location 250B may be a deep freezer in the same building. When the temperatures at the first load location 250A and the second load location 250B are different, they may be said to represent different independent thermal loads on the multiple-load vapor compression system 300.

[0053] The first load location 250A is in thermal communication with the first evaporator 240A through the first evaporator loop 245A, and the second load location 250B is in thermal communication with the second evaporator 240B through the second evaporator loop 245B. The first evaporator has a first-evaporator output temperature T_L1 that may be measured by first-evaporator sensor 242A, and the second evaporator has a second-evaporator output temperature T_L2 that may be measured by second-evaporator sensor 242B.

[0054] The control apparatus 290 of the multiple-load vapor compression system 400 may be electronically coupled to the adjustable-capacity compressor 220, the rejection apparatus 265, the first-evaporator sensor 242A, the first expansion valve 230A, the second-evaporator sensor 242B, the second expansion valve 230B, and the liquid injection valve 282 (when present).

[0055] 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 240A may be selected, and a second desired set point output temperature T_L2 for the second evaporator 240B may be selected. When more than two evaporators are present, set points for each additional evaporator may also be selected.

[0056] 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 adjustable-capacity 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 expansion valve 230A to increase the flow of refrigerant toward the first evaporator 240A. Alternatively, the control apparatus 290 may increase or decrease the cycle time and/or the speed of the adjustable-capacity compressor 220 to ensure all output temperatures (T_L1, T_L2, for example) at all evaporators 240A, 240B are within an acceptable range of their desired set-points.

[0057] 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.

[0058] 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.

[0059] 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.

[0060] 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.

[0061] 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.

[0062] 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.