Patent Description:
Heat rejection devices are widely used in many applications to exchange heat and/or provide cooling. Particular applications for heat rejection devices may include refrigeration for supermarkets, process cooling in industry, and climate control in buildings. While heat rejection is generally well understood, a number of different principles are briefly discussed for reference. A supercritical fluid (SCF) is any substance at a temperature and pressure above its critical point, where distinct liquid and gas phases do not exist. The SCF expands to fill its container like a gas but with a density similar to that of a liquid. Typically, refrigeration systems using a refrigerant with a low critical temperature utilize a transcritical cycle to reject heat to the ambient environment. The transcritical cycle is a closed thermodynamic cycle where the refrigerant goes through both subcritical and supercritical states. While many substances have a critical temperature, only substances that have a critical temperature within a useful range of temperatures and pressures are suitable for use as refrigerants.

Traditionally, chlorofluorocarbons (CFCs) have been utilized as refrigerants. However, due to the negative impact CFCs have on the environment, other refrigerant chemistry is needed.

Accordingly, it is desirable to provide an evaporative heat rejection device for refrigerants with a low critical temperature that can offer improved performance or efficiency and/or without undesirably impacting the environment, increasing the size of the unit, the manufacturing cost of the unit, and/or operating cost of the unit. <CIT> discloses a gas cooler including an air-cooled stage and an evaporative stage. The air-cooled stage provides a first cooling stage to the refrigerant discharged from the compressor, while the evaporative stage provides a second cooling stage to the refrigerant discharged from the air-cooled stage. The gas cooler cools the refrigerant below the dry bulb ambient temperature by directing the refrigerant through the air cooled stage and through the evaporative stage.

The foregoing needs are met, at least in part, by the present invention as defined by the appended claims.

An embodiment of the present invention provides an evaporatively cooled refrigeration system. The evaporatively cooled refrigeration system includes a refrigerant, a gas/liquid separator, an expansion valve in fluid connection to the gas/liquid separator, an evaporator to receive the refrigerant from the expansion valve, a compressor configured to compress the refrigerant in fluid connection to the evaporator, and a gas cooler in fluid connection to the compressor. The gas cooler includes an indirect heat exchanger to convey the refrigerant and facilitate heat from the refrigerant and a spray system to spray an evaporative coolant on the indirect heat exchanger. The gas cooler optionally includes a direct heat exchanger. Evaporative cooling provided by the evaporative coolant on the indirect heat exchanger is configured to cool the refrigerant below a dry bulb ambient air temperature.

The present invention relates to a gas cooler as defined in claim <NUM>. The gas cooler includes an indirect heat exchanger and a spray system. The indirect heat exchanger has a coil to convey a refrigerant and facilitate heat removal from the refrigerant. The spray system is designed to spray an evaporative coolant on the indirect heat exchanger. The evaporative cooling is provided by the evaporative coolant on the coil and is configured to cool the refrigerant below a dry bulb ambient air temperature.

In yet another embodiment, an evaporatively cooled refrigeration system comprises a distribution system for providing an evaporative coolant to an indirect heat exchanger. The indirect heat exchanger includes a coil configured to cool a refrigerant flowing through the coil below a dry bulb ambient air temperature by transferring heat from the refrigerant to the evaporative coolant that passes over the coil. The evaporatively cooled refrigeration system further includes a pressure relief valve configured to automatically open at a pressure of about <NUM> pounds per square inch (PSI) or greater. The indirect heat exchanger is configured to withstand a pressure of at least about <NUM> bar (<NUM> PSI).

In this respect, before explaining at least one embodiment of the invention in detail, it is to be understood that the invention is not limited in its application to the details of construction and to the arrangements of the components set forth in the following description or illustrated in the drawings. Also, it is to be understood that the phraseology and terminology employed herein, as well as the abstract, are for the purpose of description and should not be regarded as limiting.

In general, embodiments of an evaporatively cooled refrigeration system described herein refer to various forms of evaporative condensers, gas coolers, and other such heat transfer/rejection devices for use with a refrigerant having a low critical temperature. For the purposes of this disclosure, the term, "low critical temperature" refers to a substance having a critical temperature between about <NUM> and about <NUM>, or <NUM> and <NUM>. More particularly, the term, "low critical temperature" refers to a substance having a critical temperature between about <NUM> and about <NUM>, or <NUM> and <NUM>.

One or more embodiments of the evaporatively cooled refrigeration system may be designed to operate as correlated to the ambient wet-bulb temperature of the air, as opposed to the ambient dry-bulb temperature. Throughout many locations, wet-bulb temperatures may be less than <NUM> throughout the majority of the year.

The evaporatively cooled refrigeration system may be configured to operate and stay below the critical point of carbon dioxide (CO<NUM>). CO<NUM> has a low critical temperature (<NUM>) and high critical pressure (<NUM> Bar). Using ammonia as a reference, the critical temperature and critical pressure of ammonia is <NUM> and <NUM> Bar, respectively. In warm climates, the low critical temperature of CO<NUM> means the system would operate beyond the critical point using a transcritical CO<NUM> system. Transcritical operation results in a significant higher compressor energy consumption, higher compressor first cost, and less compressor life.

Furthermore, in comparison to embodiments utilizing evaporative cooling, air cooled and adiabatic cooled CO<NUM> condensing/cooling technologies require significantly greater heat rejection energy and/or much larger heat transfer surface area. It is an advantage of some embodiments that the evaporatively cooled refrigeration system is configured to operate with either condensing refrigerant vapor below its critical temperature, or cool refrigerant supercritical fluid above its critical temperature.

Moreover, CO<NUM> has a low toxicity and low environmental impact as compared with conventional refrigerants. For example, the carbon footprint of CO<NUM> is <NUM> global-warming potential (GWP). Further CO<NUM> has an Ozone depletion potential (ODP) of <NUM>. However, common refrigerants such as R-<NUM> and R-<NUM> have a GWP of <NUM>,<NUM> and <NUM>,<NUM>, and an OPD of <NUM> and <NUM>, respectively. Thus, there is a movement in the refrigeration industry to phase out the use of high GWP and ODP refrigerants.

Embodiments of the evaporatively cooled refrigeration system may include evaporative condenser coils configured to operate at internal pressures of about <NUM> Bar. The materials of the evaporative condenser coils are designed to be sufficiently corrosion-resistant to operate in a wetted environment. For example, the coils may be comprised of stainless steel, copper, galvanized steel, aluminum, or other similar materials. If the system operates in the supercritical state, the evaporative cooling technology is designed to operate at a low energy consumption compared to air cooled and adiabatic cooled systems, from both the compression and heat-rejection processes.

In some warm climates, a refrigeration system using the CO<NUM> may be configured to operate as a transcritical system. When the high side of the CO<NUM> refrigeration system operates in the supercritical state, it consumes significantly more energy from both the compression and the supercritical fluid (gas) cooling processes. By being configured to provide an evaporative cooled high side heat rejection, embodiments of the disclosure facilitate operation in the subcritical state for as many hours during a year as possible. If the evaporatively cooled refrigeration system operates in the supercritical state due to an elevated environmental temperature, for example, the evaporative cooling technology facilitates a lower system energy consumption compared to air cooled and adiabatic cooled systems.

Referring now to <FIG> of the drawings, a detailed description of the operation of an evaporatively cooled refrigeration system is disclosed. Specifically, <FIG> illustrate a simplified system schematic of an evaporatively cooled refrigeration system <NUM>. <FIG> illustrate various isometric views of a gas cooler <NUM> of <FIG> and <FIG>. <FIG> is a diagram of fluid properties plotted against pressure (ordinate) verses enthalpy (abscissa) of an evaporatively cooled refrigeration system as described in the above embodiments. By comparing <FIG> and <FIG>, the energy efficiency of the evaporatively cooled refrigeration system <NUM> can be seen.

Referring specifically to <FIG>, at step <NUM>, a flow of a refrigerant (e.g., CO<NUM>) enters a liquid/vapor separator <NUM>. The refrigerant may pass through a throttling valve <NUM>. After passing through the throttling valve <NUM>, the refrigerant may be provided in the form of a liquid/vapor mixture at an intermediate pressure, for example, about <NUM> Bar. Depending on the pressure and temperature of the gas cooler <NUM>, the liquid/vapor mixture may have a quality around <NUM>, meaning it is <NUM>% liquid, <NUM>% vapor (by mass). The liquid/vapor mixture separates into liquid and vapor in the separator <NUM> due to the difference in density of the two states.

At step <NUM>, liquid refrigerant may be drawn from the bottom of the separator <NUM>. For example, located at the bottom of the separator <NUM>, a port provides an exit for the refrigerant to be drawn from the separator <NUM> as a saturated liquid, typically around -<NUM> and <NUM> Bar.

At step <NUM>, the flow of refrigerant is provided to one or more medium temperature evaporators <NUM>. For example, the liquid refrigerant is conveyed through an expansion valve <NUM> and enters the one or more medium temperature evaporators <NUM>. Preferably, the refrigerant experiences a limited amount of expansion as it passes through the expansion valve <NUM> and enters the one or more medium temperature evaporators <NUM>. For example, the refrigerant may be nearly saturated at about -<NUM> and <NUM> Bar as it enters the one or more medium temperature evaporators <NUM>. As is generally understood, once the refrigerant enters the one or more medium temperature evaporators <NUM>, the refrigerant absorbs heat. For example, the one or more medium temperature evaporators <NUM> may be designed to have a rejection heat load of about <NUM>. <NUM> kwh ( <NUM>,<NUM> British thermal units per hour (BTU/hr)).

It is to be understood that although <FIG> shows the evaporatively cooled refrigeration system <NUM> comprising one medium temperature evaporator <NUM>, this is not to be considered limiting. The evaporatively cooled refrigeration system <NUM> may comprise more or fewer medium temperature evaporators <NUM> depending on the embodiment. For example, for higher heat load systems, more than one medium temperature evaporator <NUM> may be needed.

At step <NUM>, the refrigerant exits the one or more medium temperature evaporators <NUM>. Having gained heat in the one or more medium temperature evaporators <NUM>, the refrigerant is provided in the form of a vapor having a small amount of superheat of less than -<NUM> (<NUM> °F) and at a pressure of about <NUM> Bar.

At step <NUM>, the refrigerant may be conveyed to one or more low temperature evaporators <NUM>. For example, the liquid refrigerant may be conveyed from the separator <NUM> to a low temperature expansion value <NUM>. The flow of refrigerant is supplied to the one or more low temperature evaporators <NUM> at about -<NUM> and <NUM> Bar. As is generally understood, once the refrigerant enters the one or more low temperature evaporators <NUM>, the refrigerant absorbs heat. For example, the one or more low temperature evaporators <NUM> may have a rejection heat load of about <NUM> kwh (<NUM>,<NUM> BTU/hr).

It is to be understood that although <FIG> shows the evaporatively cooled refrigeration system <NUM> comprising one low temperature evaporator <NUM>, this is not to be considered limiting. The evaporatively cooled refrigeration system <NUM> may comprise more or fewer low temperature evaporators <NUM> depending on the embodiment. For example, for higher heat load systems, more than one low temperature evaporator <NUM> may be needed.

It is to be further understood that although <FIG> shows the evaporatively cooled refrigeration system <NUM> as a two-stage system (i.e., having one or more medium temperature evaporators <NUM> and one or more low temperature evaporators <NUM>), this is not to be considered limiting. In one embodiment, the evaporatively cooled refrigeration system <NUM> may be a single-stage system having either the medium temperature evaporator <NUM> or the low temperature evaporator <NUM>.

At step <NUM>, the refrigerant exits the one or more low temperature evaporators <NUM>. For example, having gained heat in the one or more low temperature evaporators <NUM>, the refrigerant is provided in the form of a vapor with a small amount of superheat, typically less than <NUM>°F, and at a pressure of about <NUM> Bar. This flow of refrigerant exiting the one or more low temperature evaporators <NUM> provides a suction for one or more low temperature compressors <NUM> provided downstream of the low temperature evaporators. The flow of refrigerant is designed to be compressed by the one or more low temperature compressors <NUM> as described in more detail hereinbelow.

Referring again to the liquid/vapor separator <NUM>, at step <NUM>, a flash gas (e.g., excess vapor that has boiled off the liquid refrigerant in the separator <NUM>) is removed from the liquid/vapor separator <NUM>. A flash gas bypass valve <NUM> may be provided and is configured to meter a flow of the flash gas from the separator <NUM>. In a particular example, the metered flow of the flash gas may be a saturated vapor provided at a pressure of about <NUM> Bar. The metered flow of the flash gas is then combined with an amount of compressed flow of refrigerant that is exiting the one or more low temperature compressors <NUM>, and the combined stream is conveyed to one or more high stage compressors <NUM>. The combined stream is designed to control or modulate the pressure of the separator <NUM> and the superheat supplied to the suction of the one or more high stage compressor(s) <NUM>.

At step <NUM>, the compressed flow of refrigerant exits the one or more low temperature compressors <NUM>. For example, the one or more low temperature compressors <NUM> may have compressed the refrigerant to slightly above the high stage suction pressure, about <NUM> Bar. The refrigerant also has superheat leaving the one or more low temperature compressors <NUM> at around <NUM>. The number of the one or more low temperature compressors <NUM> may depend on the needs of the embodiment. Thus, the amount of superheat leaving the one or more low temperature compressors <NUM> can vary.

At step <NUM>, the flow of refrigerant is provided to one or more high stage compressors <NUM>. In this instance, a high stage compressor head serves to converge the flow of refrigerant from the outlet of the low temperature compressor(s) <NUM>, the outlet of the medium temperature evaporator(s) <NUM>, and the flash gas bypass valve <NUM>. The converged flow of refrigerant is provided to the one or more high stage compressors <NUM> and compressed.

It is to be understood that the evaporatively cooled refrigeration system <NUM> depicted herein is designed to operate at a low pressure during steps <NUM> - <NUM>. Thus, the piping and one or more components of the portion of the evaporatively cooled refrigeration system <NUM> on the "low side" may be rated for low pressure environments. One of the benefits of using CO<NUM> as a refrigerant is that the piping on the low side may be smaller than conventional refrigeration systems. For example, at -<NUM> (-<NUM> °F), R-<NUM> has a vapor density of about <NUM>/m<NUM> (<NUM> pounds per cubic foot (lbs/ft<NUM> )). Whereas, at the same temperature, CO<NUM> has a vapor density of about <NUM>/m<NUM> (<NUM> lbs/ft<NUM>). Thus, the same mass flow rate of CO<NUM> can be transported through a smaller pipe as compared to R-<NUM>.

Referring again to <FIG>, at step <NUM>, the flow of refrigerant is discharged from the high stage compressor(s) <NUM>, where the flow of refrigerant then enters the gas cooler <NUM>, which is described in further detail below. In this instance, the refrigerant has been compressed and is now at the highest temperature and pressure in the evaporatively cooled refrigeration system <NUM>. For example, Table <NUM> shows the pipe schedule required for various nominal pipe sizes for the "high side" of the evaporatively cooled refrigeration system <NUM>.

At an inlet <NUM> of an indirect heat exchanger <NUM> of the gas cooler <NUM>, as further explained below, the coupling between the inlet <NUM> and the pipe leading from the high stage compressor <NUM> must also be able to withstand the high pressure and temperature described above. Further, the highpressure portion of the evaporatively cooled refrigeration system <NUM> may comprise one or more safety relief valves configured to automatically vent at about <NUM> bar ( <NUM> psi). Accordingly, a pipe coupling <NUM> between the inlet <NUM> of the indirect heat exchanger <NUM> and the pipe leading from the high stage compressor <NUM> must be capable of withstanding about <NUM> bar ( <NUM> psi) without separating or leaking.

The discharge of the high stage compressor <NUM> may pass through an oil separator <NUM>. The refrigerant that is discharged from the high stage compressor(s) <NUM> is also provided to the inlet <NUM> of the gas cooler <NUM>. The gas cooler <NUM> may be provided in the form of a gas condenser. The gas cooler <NUM> is configured to provide evaporative cooling. In some forms, an adiabatic cooler is not provided in the evaporatively cooled refrigeration system <NUM>. In this manner, efficient cooling/condensing of refrigerants with low critical temperatures may be achieved. Primarily the throttling valve <NUM>, along with capacity modulation of the gas cooler <NUM>, controls the pressure at this point while the refrigerant is passing through the cooler. During transcritical operation, when the ambient dry bulb temperature, or adiabatically pre-cooled air dry bulb temperature is greater than the CO<NUM> critical temperature (<NUM>), an air-cooled or adiabatically cooled refrigeration system (as compared to the evaporatively cooled refrigeration system <NUM> shown in <FIG>) operates at the supercritical region of the pressure verses enthalpy diagram illustrated in <FIG>. The temperature of the CO<NUM> may reach between <NUM> and <NUM> and <NUM> Bar to103 Bar. The compression process consumes a significant amount of energy and thus adds a significant amount of additional waste heat to be rejected to the ambient atmosphere from the low stage compression process (steps <NUM>-<NUM>).

As best seen in <FIG>, at step <NUM>', during subcritical operation, the refrigerant is evaporatively cooled. In subcritical operation, the ambient heat sink is with reference to the wet bulb temperature.

As shown in <FIG>, the gas cooler <NUM> for use in the systems described herein may include a housing <NUM> that surrounds the internal components of the gas cooler <NUM>. Referring to <FIG>, and <FIG>, the gas cooler <NUM> may include a first distribution system <NUM> designed to supply an evaporative fluid, such as water, water solution, or the like to a direct heat exchanger <NUM>. In a simplified example, water may be provided to and directly sprayed onto the direct heat exchanger <NUM>. The first distribution system <NUM> may further include a hot water basin <NUM> and a first plurality of nozzles <NUM> configured to distribute water to the direct heat exchanger <NUM>. In use, water may cascade down sheets of fill media disposed in the direct heat exchanger <NUM>. An air movement device <NUM>, such as a fan, draws a flow of air through the fill media to evaporatively cool the evaporative fluid. The flow of air may then enter the direct heat exchanger <NUM> through an upper air inlet <NUM>. In various examples, the air movement device <NUM> may include more than one fan and the size may vary depending upon the size of the gas cooler <NUM> and specific application. Thus, the direct heat exchanger <NUM> is designed to cool the evaporative fluid from a first temperature to a second temperature less than the first temperature.

In some embodiments, the cooled evaporative fluid may be collected in a second distribution system <NUM> comprising an intermediate basin <NUM> and a second plurality of nozzles <NUM> configured to spray the evaporative fluid onto the indirect heat exchanger <NUM>. However, in some other embodiments, the evaporatively cooled refrigeration system <NUM> may not include the intermediate basin <NUM>. Thus, the evaporative fluid may pass through the direct heat exchanger <NUM> and then be provided for use with the indirect heat exchanger <NUM> without flowing through a secondary distribution system. In some examples, the indirect heat exchanger <NUM> may include one or more cooling/condensing coils, a plate heat exchanger, or the like. The indirect heat exchanger <NUM> may further include a drift eliminator <NUM>. The indirect heat exchanger <NUM> is configured to cool the refrigerant flowing through the tube side of the indirect heat exchanger <NUM> by transferring heat from the refrigerant to the evaporative fluid. Thus, the evaporative fluid increases in temperature after contacting the indirect heat exchanger <NUM>. The refrigerant may be provided to and enter the indirect heat exchanger <NUM> at the inlet <NUM> and exit the indirect heat exchanger <NUM> at an outlet <NUM>. To aid in the transfer of heat via the indirect heat exchanger <NUM>, the air movement device <NUM> may draw air in from a lower air inlet <NUM> and through the indirect heat exchanger <NUM>.

A collection basin <NUM> may be provided to collect the hot evaporative fluid that has passed over the indirect heat exchanger <NUM>. A recirculation system may then convey the hot evaporative fluid from the collection basin <NUM> to the hot water basin <NUM> in a well understood manner. In one non-limiting example, the recirculation system includes a pump <NUM> and piping <NUM>. In this way, the evaporative fluid may be cycled over the gas cooler <NUM>. However, due to evaporation, a supply of fresh water may be supplied to the gas cooler <NUM> to maintain a desired evaporative fluid volume. For example, the evaporative fluid may be cycled through the evaporatively cooled refrigeration system <NUM> at least three times before the evaporative fluid may need to be removed from the evaporatively cooled refrigeration system <NUM> through blowdown.

In most climates, the ambient wet bulb is <NUM> or lower. The highest temperature and pressure point of the evaporatively cooled refrigeration system <NUM> hence would be able to stay below the CO<NUM> critical point. Referring again to <FIG>, instead of a transcritical compression process from <NUM> to <NUM>, the evaporatively cooled refrigeration system <NUM> is designed to operate at a subcritical compression process from <NUM> to <NUM>'. The line lengths of <NUM>-<NUM> and <NUM>'-<NUM>' is proportional to the energy of the waste heat to be rejected from both the low stage and high stage compression processes. Since the line length of <NUM>'-<NUM>' is shorter than <NUM>-<NUM>, it represents a reduced total compression energy consumed. It is an advantage of the evaporatively cooled refrigeration system <NUM> that the gas cooler <NUM> is configured to utilize evaporative cooling which increases a percentage of time the evaporatively cooled refrigeration system <NUM> operates subcritically in comparison to a system that lacks evaporative cooling. In this manner, the evaporatively cooled refrigeration system <NUM> minimizes the pressure required and associated energy consumption. Even in rare climates (or extreme hot hours) in which the ambient wet bulb is higher than the CO<NUM> critical temperature, the evaporatively cooled refrigeration system <NUM> facilitates a transcritical compression line below line <NUM>-<NUM> which indicates a lower pressure and temperature compression compared line <NUM>-<NUM>.

As shown in <FIG>, at step <NUM>, heat has been removed from the refrigerant via the gas cooler <NUM> and the refrigerant exits the gas cooler <NUM>. In transcritical mode, the refrigerant at this point is provided as an undefined fluid, as opposed to a liquid or a gas.

In contrast, at step <NUM>', the refrigerant exits the gas cooler <NUM> as a liquid. In subcritical mode, the refrigerant at this point is a liquid and heat has been removed from the refrigerant via the gas cooler <NUM>.

At step <NUM>, the refrigerant cycle is completed as the refrigerant exits the outlet of the throttling valve <NUM>. When passing through the throttling valve <NUM>, the flow of refrigerant from the gas cooler <NUM> has been reduced in pressure to the pressure substantially equivalent to, or equal to the pressure within the liquid/vapor separator <NUM>. In transcritical mode, the refrigerant here is a mixture of liquid and gas. In subcritical mode, the refrigerant is <NUM>% liquid. Thus, piping downstream of the throttling valve <NUM> may be rated for lower pressure operation.

<FIG> is a diagram of fluid properties of carbon dioxide (CO<NUM>) plotted against pressure (ordinate) verses temperature (abscissa) suitable for use as the refrigerant in the evaporatively cooled refrigeration system <NUM>. As shown in <FIG>, the critical point of carbon dioxide is about <NUM>°K (<NUM>) and about <NUM> Bar.

In many climates, the annual operating cost of the evaporatively cooled refrigeration system <NUM> is lower than the annual operating cost of a conventional adiabatic cooled refrigeration system with respect to the cost of water usage and energy usage in the instance where each of the evaporatively cooled refrigeration system <NUM> and the conventional adiabatic cooled refrigeration system provide substantially the same cooling capacity. For example, for a climate C1 with a mean coincident wet bulb temperature of MCWB1, a dry bulb temperature of DB1, and a wet bulb temperature of WB1, having water cost WC1 USD/Gal and energy cost EC1 USD/kWh, the conventional adiabatic cooled refrigeration system has a total annual cost of TACA1 to provide X BTU output. In contrast, in the same climate C1 and providing the same X BTU output, the evaporatively cooled refrigeration system <NUM> has a total annual cost of TACE1 which is less than TACA1. In C1, the annual energy use AEU1 of the evaporatively cooled refrigeration system <NUM> is less than the annual energy use AEU2 of the conventional adiabatic cooled refrigeration system. In some climates, the annual water use AWU1 of the evaporatively cooled refrigeration system <NUM> is higher than the annual water use AWU2 of the conventional adiabatic cooled refrigeration system. However, AEU1 x EC1 + AWU1 x WC1 is less than AEU2 x EC1 + AWU2 x WC1, and, therefore, TACE1 is less than TACA1. Accordingly, the cost savings in annual energy use of the evaporatively cooled refrigeration system <NUM> outweighs the cost of any additional water use. This is found to be true for a wide range of energy costs EC1, water costs WC1, and climates C1.

In <FIG>, a few examples are provided that illustrate the differences between a conventional adiabatic cooled refrigeration system and the evaporatively cooled refrigeration system <NUM> to amplify and elaborate on the foregoing concepts.

Turning to <FIG>, a Table <NUM> comprising simulated operating parameters for a conventional adiabatic system is shown. The table includes values for an adiabatic system operated in three different simulated geographical environments - California, New York, and North Dakota. Each of the three locations have different average climate conditions (e.g., weather) such as temperature and humidity. California is simulated as having a mean coincident wet bulb temperature of <NUM> (<NUM> °F), a dry bulb temperature of <NUM> (<NUM> °F) and a wet bulb temperature of <NUM> (<NUM> °F). New York is simulated as having a mean coincident wet bulb temperature of <NUM> (<NUM> °F), a dry bulb temperature of <NUM> (<NUM> °F), and a wet bulb temperature of <NUM> (<NUM> °F). North Dakota is simulated as having a mean coincident wet bulb temperature of <NUM> (<NUM> °F), a dry bulb temperature of <NUM> (<NUM> °F), and a wet bulb temperature of of <NUM> (<NUM> °F).

<FIG> shows a Table <NUM> that comprises simulated operating parameters for an evaporative cooler system, such as the evaporatively cooled refrigeration system <NUM> of <FIG>. The Table <NUM> displays the operating parameters of the evaporatively cooled refrigeration system <NUM> in the same simulated environments provided for the adiabatic system, e.g. California, New York, and North Dakota, with the above-referenced climate characteristics.

Each of the adiabatic and the evaporative cooler systems of <FIG> and <FIG> are designed to have a cooling capacity of about <NUM> kwh ( <NUM>,<NUM> BTUs) to about <NUM> kwh (<NUM>,<NUM> BTUs). An Annual Operation Cost for each system was calculated for the three different geographical locations. The Annual Operation Cost may be a function of water cost (CW) and energy cost (CE).

By comparing the values in Table <NUM> to Table <NUM>, it can be seen that the evaporative cooler system has a lower Annual Operation Cost compared to the adiabatic system. As shown in Table <NUM>, the evaporative cooler system may have a lower annual operation cost because the evaporative cooler may have a lower power requirement for a fan, such as the air movement device <NUM> of <FIG>. Thus, less power may be needed; thereby reducing cost. Further, the evaporative cooler may be able to operate at higher cycles of concentration of the evaporative coolant fluid which may reduce cost. Moreover, the evaporative refrigeration system may achieve the same cooling capacity as the adiabatic cooler system but also may be provided with a footprint (e.g., size of the housing and/or system) and/or overall size that is between about <NUM> to about <NUM> percent of the size of the adiabatic cooler. In some forms, the conventional adiabatic cooled refrigeration system has a footprint X<NUM> and the evaporatively cooled refrigeration system <NUM> has a footprint Y<NUM>. In some forms, despite providing substantially the same cooling capacity, the ratio of X<NUM> to Y<NUM> is between about <NUM>:<NUM> to about <NUM>:<NUM>. By having a smaller footprint, the amount of materials needed to construct the system may be reduced and the available land needed may be smaller; thereby saving on material and land cost. Thus, not using an adiabatic cooling system, and instead using the evaporatively cooled refrigeration system <NUM> saves both money and space.

A summary of the Overall Annual Cost of the Adiabatic System of <FIG> and the Evaporative System of <FIG> is shown in Table <NUM> of <FIG>. As can be seen, in a variety of environments, the evaporative cooler system has a lower Overall Annual Operation Cost even when the evaporative cooler system operates at higher water usage rates because the energy requirements of the evaporative cooler system is lower than that of a conventional adiabatic system.

Referring back to <FIG>, the evaporatively cooled refrigeration system <NUM> may be connected to a controller <NUM>. The controller <NUM> may be connected to the evaporatively cooled refrigeration system <NUM> over a wireless network or a cable network. The controller <NUM> may be configured to receive information from and/or send commands to the evaporatively cooled refrigeration system <NUM>. For example, the controller <NUM> may be configured to determine the ambient air temperature and the required heat load of the evaporatively cooled refrigeration system <NUM>. Using this information, the controller <NUM> may adjust one or more components of the evaporatively cooled refrigeration system <NUM>.

In one scenario, the controller <NUM> may adjust the fan speed of the air movement device <NUM> of <FIG> and <FIG> to achieve the lowest overall refrigeration system energy. When the ambient air temperature is colder, it may not be necessary to operate the air movement device <NUM> at a full rate. Thus, the fan speed may be reduced, thereby reducing energy cost.

In yet another scenario, the controller <NUM> may adjust the speed of the low temperature compressor <NUM> and/or the high stage compressor <NUM>. For example, the evaporatively cooled refrigeration system <NUM> may have a low heat load, which does not require one or more of the compressors <NUM>, <NUM> to operate at full speed. Thus, the controller <NUM> may be configured to reduce the speed of the low temperature compressor <NUM> and/or the high stage compressor <NUM>. Therefore, the energy needs of the evaporatively cooled refrigeration system <NUM> may be reduced.

It is to be understood that the above examples are merely for illustrative purposes and are not to be considered limiting. The controller <NUM> may be configured to adjust multiple components of the evaporatively cooled refrigeration system <NUM> simultaneously.

Additionally, it is to be understood that carbon dioxide is one example of a suitable refrigerant for use in the evaporatively cooled refrigeration system <NUM>. Other refrigerants having similar properties to that of carbon dioxide may be used. For example, the evaporatively cooled refrigeration system <NUM> may use refrigerants having one or more of the following properties: a low ODP (near zero), a low GWP, a low critical temperature (less than <NUM>), and a high critical pressure (greater than <NUM> bar (<NUM> psig)).

Claim 1:
A gas cooler (<NUM>), comprising:
an indirect heat exchanger (<NUM>) comprising a coil;
a distribution system (<NUM>) for providing an evaporative coolant to the indirect heat exchanger (<NUM>); and
a direct heat exchanger (<NUM>) configured to cool the evaporative coolant from a first temperature to a second temperature less than the first temperature prior to the evaporative coolant being provided to the indirect heat exchanger (<NUM>),
wherein the indirect heat exchanger (<NUM>) is configured to cool a refrigerant flowing through the coil below a dry bulb ambient air temperature by transferring heat from the refrigerant to the evaporative coolant provided to the indirect heat exchanger (<NUM>).