Active/passive cooling system

A cooling assembly includes an evaporator containing a primary cooling medium, a passive condenser, and a heat exchanger. When a secondary cooling medium is provided to the heat exchanger, the primary cooling medium in the gas phase switches from being received by the passive condenser to the heat exchanger without operating any valves located between the evaporator and the passive condenser and between the evaporator and the heat exchanger. The primary cooling medium circulates between the evaporator and the passive condenser and between the evaporator and the heat exchanger by natural circulation and gravity without a pump in the flow path of the primary cooling medium between the heat exchanger and the evaporator and between the passive condenser and the evaporator to circulate the primary cooling medium.

FIELD OF THE INVENTION

This invention relates to cooling systems and systems and methods to control them. In particular, this invention relates to a cooling system having both active and passive modes. A particularly suitable application, for example, is in data center cooling systems.

BACKGROUND OF THE INVENTION

Data centers often require large amounts of energy to operate. The servers in these data centers generate a large amount a heat, requiring cooling. To reduce the energy use of data centers, more efficient cooling systems are desired.

Heat pipes and thermosyphons are devices that transfer energy from a higher temperature evaporator section to a lower temperature condenser section by the evaporation and condensation of a closed refrigerant volume. Transfer of the refrigerant from the condenser section to the evaporator section takes place either by gravity or capillary force. Heat pipes have been used in data center cooling as indirect economizers. In these installations, warm air from a data center, for example, is recirculated through the evaporator section of a heat pipe where the enclosed refrigerant is vaporized by the heat from the data center, cooling the data center air. Cooler ambient air is blown over the condenser section of the heat pipe where the refrigerant vapor is condensed and the data center heat expelled. In some applications, the ambient air is first adiabatically cooled with an evaporative cooler prior to its passage through the condenser section of the heat pipe to provide a lower temperature heat sink. In another configuration, the condenser section of the heat pipe may be sprayed with water at the same time as the ambient air passes across its surface, providing a heat sink temperature close to the ambient wet bulb temperature.

In these implementations, the heat pipes and thermosyphons are limited by the ambient temperature conditions and may not provide sufficient heat rejection when the ambient temperatures are high. One such solution to this limitation is a pumped refrigerant system that incorporates a mechanical cooling system, such as a direct expansion (DX) cooling system (an active mode), with a near passive mode that operates similarly to a thermosyphon. These systems include a pump to move liquid refrigerant from the condenser to the evaporator. By utilizing a pump, the flow of refrigerant can be controlled independently of the evaporator and condenser pressure drops and the effects of gravity. This approach is near passive because only a small amount of power is needed for the pump to transfer a significant amount of thermal energy. In the pumped refrigerant systems, the pump can be turned off and, by operating valves, compressors and expansion valves can be integrated into the system refrigerant flow to allow the system to act as a direct expansion cooling system.

System design constraints in direct expansion systems generally call for modest pressure drops in the evaporator and condenser section of the system to provide even refrigerant flow through the multiple parallel evaporator and condenser circuits of the system. It is because of these pressure drops that the system requires a pump to circulate the refrigerant fluid when running in the near passive mode. As the refrigerant circuit includes a compressor system, the refrigerant volume also needs to include oil for lubrication. There are various design and operational constraints, such as best practices for refrigerant velocity, (so called “oil management”) for the pumped refrigerant systems that ensure oil does not get trapped in the various piping lengths of the system and returns reliably to the compressor where it is needed. These oil management constraints become problematic when operating in the pumped (near passive mode) because the flow paths and flow rates for the pumped mode may not coincide with the rules required for the DX mode. Refrigerant volume for proper operation within the system may also be quite different during the pumped mode and the DX mode due to differing available superheat and sub-cool levels, and coil flooding levels, for example.

Cooling systems with further reductions in energy use are thus desired, as are cooling systems that do not require oil management in a passive or near passive mode.

SUMMARY OF THE INVENTION

In one aspect, the present invention relates to a cooling system including an evaporator containing a primary cooling medium, a passive condenser, and a heat exchanger. The evaporator is configured to receive a process fluid and, when receiving the process fluid, to change the phase of the primary cooling medium from liquid to gas. The passive condenser has an outer surface and is fluidly coupled to the evaporator. The passive condenser is configured to have an airstream directed over the outer surface thereof. When the airstream is directed over the outer surface of the passive condenser, the passive condenser is configured (i) to receive the primary cooling medium in the gas phase from the evaporator, (ii) to transfer heat from the primary cooling medium, (iii) to change the phase of the primary cooling medium from gas to liquid, and (iv) to supply the primary cooling medium in the liquid phase to the evaporator. The heat exchanger is fluidly coupled to the evaporator and configured to have a secondary cooling medium selectively provided thereto. When the secondary cooling medium is provided to the heat exchanger, the primary cooling medium in the gas phase switches from being received by the passive condenser to the heat exchanger without operating any valves located between the evaporator and the passive condenser and between the evaporator and the heat exchanger. The heat exchanger is configured (i) to receive the primary cooling medium in the gas phase from the evaporator, (ii) to transfer heat from the primary cooling medium, (iii) to change the phase of the primary cooling medium from gas to liquid, and (iv) to supply the primary cooling medium in the liquid phase to the evaporator. When the heat exchanger is not accepting the secondary cooling medium, the heat exchanger does not supply the primary cooling medium in the liquid phase to the evaporator. The primary cooling medium circulates between the evaporator and the passive condenser and between the evaporator and the heat exchanger by natural circulation and gravity without a pump in the flow path of the primary cooling medium between the heat exchanger and the evaporator and between the passive condenser and the evaporator to circulate the primary cooling medium.

In another aspect, the invention relates to a method of cooling a process fluid. The method includes directing a process fluid through an evaporator to transfer heat from the process fluid to a primary cooling medium contained in the evaporator and change the primary cooling medium from a liquid phase to a gas phase and selectively utilizing one of a heat exchanger and a passive condenser to change the primary cooling medium from the gas phase to the liquid phase. Each of the heat exchanger and the passive condenser are coupled to the evaporator to receive the primary cooling medium in the gas phase from the evaporator and supply the primary cooling medium in the liquid phase to the evaporator. The method also includes circulating the primary cooling medium between the evaporator and at least one of the heat exchanger and the passive condenser by natural circulation. The one of the heat exchanger and the passive condenser is selected without operating any valves in the coupling between the evaporator and the passive condenser and in the coupling between the evaporator and the heat exchanger.

DETAILED DESCRIPTION OF THE PREFERRED EMBODIMENTS

FIG.1shows a data center100having a cooling system110according to a preferred embodiment of the invention.FIG.2is a cross-section view of the cooling system110taken along line2-2inFIG.1. Although the cooling system110is shown and described in reference to a data center100, the cooling system110is not limited to this application and may be used in other suitable air cooling applications. Electronic components such as servers may be mounted on racks102, and in a data center100, these racks102may be arranged in rows forming aisles104,106, therebetween. One aisle104is a cold aisle, and another aisle106is a hot aisle. Cool, supply air112from the cooling system is directed into the cold aisle104. The supply air112then passes from the cold aisle104through the racks and into the hot aisle106. As the air passes through the racks102, it draws heat from the electronic components, cooling them and resulting in hot air passing into the hot aisle106. This air is then directed back to the cooling system110as hot, return air114. Supply air fans116are used to draw the return air114from the data center100, pass the return air114through the cooling system110, where it is cooled, and then return the now cooled return air114to the data center100as supply air112. The portion of the cooling system110through which the return air114flows, is cooled, and is returned as supply air112is referred to herein as the interior air handler132.

The cooling system110uses at least one airstream cooling assembly to cool the return air114. The airstream cooling assemblies described in the following embodiments may also be referred to as an airstream cooling assembly loop or a loop.FIGS.3and4show an airstream cooling assembly200according to a first embodiment of the invention. The airstream cooling assembly200has two modes, a passive mode and an active mode. The passive mode may also be referred to as economization mode.FIG.3is a schematic of the airstream cooling assembly200in the passive mode, andFIG.4is a schematic airstream cooling assembly200in the active mode. The airstream cooling assembly200incorporates the efficiency of a thermosyphon with the ability to provide active cooling when available ambient free cooling sinks are not at a low enough temperature to provide sufficient heat rejection. This is accomplished by including two separate condensers214,216in the loop200, one of which (condenser214) is used in the passive mode and the other of which (condenser216) is used in the active mode.

The airstream cooling assembly200circulates a primary cooling medium202through a primary coolant loop210. The primary cooling medium202circulates through the primary coolant loop210by natural circulation and gravity without the need for pumps and compressors. The primary cooling medium202may be any suitable refrigerant that changes phase from a liquid to a gas. As will be discussed further below, the primary coolant loop210does not require any moving parts. As a result, the available range of refrigerants that are suitable as primary cooling mediums202is greatly expanded compared to direct expansion (DX) cooling systems, for example, and suitable refrigerants include natural refrigerants such as water.

The primary coolant loop210includes an evaporator212, and the primary cooling medium202is contained within the evaporator212. In this embodiment, the evaporator212is a coil and preferably a one-pass, flooded coil. Any suitable coil may be used including, for example, microchannel coils, such as those described further below, or finned tube coils. In both the passive mode and the active mode, the return air114is directed over the outer surface of the evaporator212by the supply air fans116. The hot, return air114evaporates the primary cooling medium202in the evaporator212as it passes over the outer surface of the evaporator212. The phase change of the primary cooling medium202from a liquid phase204to a gas (or vapor) phase206cools the return air114, allowing it to be returned to the data center100as cool, supply air112. The vapor206then rises through a vapor pipe222to one of the two condensers214,216.

In the passive mode, shown inFIG.3, the vapor206travels to a passive condenser214in the primary coolant loop210. As with the evaporator212, the passive condenser214of this embodiment is a coil, preferably a one-pass coil, and any suitable coil may be used including, for example, microchannel coils, such as those described further below, or tube coils (both finned and unfinned). Scavenger air118is drawn across an outer surface of the passive condenser214by scavenger fans120(seeFIGS.1and2). In this embodiment, the scavenger air118is ambient air drawn from the outdoor environment surrounding the cooling system110. As the scavenger air118passes over the passive condenser214, the heat of the primary cooling medium202contained in the passive condenser214is released to the scavenger air118, condensing the vapor206to a liquid204. Gravity then causes the primary cooling medium202, now in the liquid phase204, to flow down a liquid refrigerant line224and to return to the evaporator212. The scavenger air118is exhausted to the outside by the scavenger fans120.

When the ambient air conditions are not sufficient to cool the return air114to the desired conditions (e.g., temperature) for the supply air112, the airstream cooling assembly200may be operated in an active mode shown inFIG.4. In the active mode, the vapor206of the primary cooling medium202is condensed in an active condenser216. In this embodiment, the active condenser216may also be referred to herein as a heat exchanger (HX). In the active condenser216, heat is transferred from the primary cooling medium202to a secondary cooling medium208of a secondary cooling system230. The secondary cooling medium208may be any suitable refrigerant medium, including, for example, cooled (or chilled) water or a vapor change refrigerant used in a direct expansion cooling system. The active condenser216may be any suitable heat exchanger including, for example, a plate heat exchanger, a coaxial heat exchanger, or a shell and tube heat exchanger. As the heat of the primary cooling medium202is rejected to the secondary cooling medium208, the primary cooling medium202condenses from a vapor206to a liquid204. As with the passive condenser214, gravity then causes the primary cooling medium202, now in liquid phase204, to flow down a liquid refrigerant line224and to return to the evaporator212.

In this embodiment, the secondary cooling system230is a direct expansion (DX) cooling system230using the common refrigeration cycle, and the secondary cooling medium208is any suitable refrigerant used in such systems. The direct expansion cooling system230includes a compressor232to increase the pressure and temperature of the refrigerant208before it is cooled in a condenser234. In this embodiment, the condenser234of the direct expansion cooling system230may also be cooled by the scavenger air118(seeFIGS.1and2). The refrigerant208then passes through an expansion valve236, reducing its pressure and temperature, before flowing into the active condenser216.

Even in the active mode, the airstream cooling assembly200operates without the need of pumps, oils, or compressors in the primary coolant loop210. The airstream cooling assembly200even operates without valves to switch between modes. Instead, the vapor206of the primary cooling medium202naturally travels to the colder of the two condensers214,216to condense. Thus, by activating the secondary cooling system230to cool the active condenser216, the airstream cooling assembly200automatically switches from passive mode to active mode (assuming the temperature in the active condenser216is lower than in the passive condenser214) and by deactivating the secondary cooling system230, the loop200searches back to the passive mode. As described below, a controller240may be used to activate and deactivate the secondary cooling system230. Another advantage of the airstream cooling assembly200attributable to the lack of moving parts is that no oil is required, allowing the primary cooling medium202to flow outside of refrigerant velocities commonly required to entrain and keep oil circulating within the primary loop210.

Although the condensers214,216of the airstream cooling assembly200are shown in parallel inFIGS.3and4, the condensers214,216may also be arranged in series so the outlet of one of the active or passive condensers214,216is upstream of and feeds into the inlet of the other condenser214,216. The primary coolant loop210may also include traps218and/or check valves220after each of the condensers214,216. The traps218and check valves220prevent reverse flow of the primary cooling medium202through the condenser214,216that is not currently operating in a given mode. Other suitable valves or methods may be used to prevent reverse flow. Such valves or traps are optional as including the traps218and check valves220may increase the pressure drop of the system and thus inhibit the natural circulation flow of the primary cooling medium202in the primary loop210.

Vent lines226may be located after the traps218of each condenser214,216and connect to the inlet of the respective condenser214,216. These vent lines226allow any gas entrained in the liquid204of the primary cooling medium202to escape to the vapor side of the loop and thus assist the liquid flow by gravity to the evaporator212.

Bubbles that form during evaporation of the primary cooling medium202in the evaporator212may entrain liquid as they rise in the channels of the evaporator212. An entrained liquid return line228may be located at the outlet of the evaporator212and connect to the inlet header of the evaporator212, allowing this entrained liquid to return to the evaporator inlet header without having to flow counter to the boiling flow path in the evaporator212.

Because the airstream cooling assembly200of this embodiment operates in natural circulation with the assistance of gravity, the evaporator212is placed at a level lower than either condenser214,216, to allow gravity to assist in returning the condensed primary cooling medium202(liquid204) to the evaporator212. Maintaining the primary cooling medium202in the liquid phase204throughout the entire length of the evaporator212is desirable. The height of the condensers214,216above the evaporator are thus preferably high enough to provide sufficient pressure head from the primary cooling medium202in the liquid phase204to overcome the pressure drop of the evaporator212. Although the evaporator212may be level with the liquid and vapor headers of the evaporator212lying in the same horizontal plane, the evaporator212may also be preferably inclined at an angle α from horizontal with the vapor header of the evaporator212higher than the liquid header to facilitate vapor expulsion. The passive condenser214may also be preferably inclined at an angle β from horizontal with the liquid header of the passive condenser214lower than the vapor header to facilitate condensate flow via gravity. The inclination angles (angles β) of the passive condenser is preferably sufficient to provide a clear drain path and eliminate back flow of the primary cooling medium202in the passive condenser214.

As discussed above, the evaporator212and passive condenser214may be microchannel coils.FIG.5Ashows a microchannel coil300that may be used as the evaporator212and passive condenser214of this embodiment. Using a microchannel coil300has a number of advantages, including, for example, that the high internal surface area of the microchannel coil300aids in heat transfer. In addition, a microchannel coil300greatly reduces the volume of the primary cooling medium202needed in the primary coolant loop210as compared to, for example, finned tube coils. This reduction in primary cooling medium202volume is beneficial for a number of reasons including reduced costs and, when certain refrigerants are used, a reduced source of potential greenhouse gas emissions. The microchannel coil300has a liquid side302and a vapor side304. As shown inFIG.5A, when the microchannel coil300is used as an evaporator212, the flow of the primary cooling medium202is from the liquid side302to the vapor side304(left to right), and when the microchannel coil300is used as a passive condenser214, the flow is opposite (right to left).

The microchannel coil300has a liquid header310and a vapor header320that are connected by a plurality of microchannel extrusions330. A cross-section, taken along line5B-5B inFIG.5A, of a microchannel extrusion is shown inFIG.5B. The microchannel extrusions330have an outer surface332and include a plurality of microchannels334,336. An airstream is directed over the outer surface332of the microchannel extrusions330in direction A, shown inFIG.5B(in and out of the page inFIG.5A). The primary cooling medium202flows through the microchannels334,336. Each of the plurality of microchannel extrusions330are mechanically brazed to aluminum fins340positioned between the microchannel extrusions330to promote heat transfer.

The liquid header310includes a liquid connection312that connects the liquid header310to the liquid refrigerant line224. Likewise, the vapor header320also includes at least one vapor connection322that connects the vapor header320to the vapor pipe222. With a microchannel coil used as the evaporator212, it may be beneficial to have multiple vapor connections322. In this embodiment, three vapor connections322are shown. Using multiple vapor connections322reduces the back pressure of vapor in the vapor header320and promotes natural circulation flow in the primary coolant loop210. When multiple vapor connections322are used for the evaporator212, a corresponding number of vapor connections322may be used for both the passive and active condensers214,216, resulting in a plurality of vapor pipes222connecting the vapor connections322. Another consideration for the vapor pipe222and the vapor connections322is to use large diameter pipes, decreasing back pressure of vapor and promoting natural circulation flow in the primary coolant loop210. For example, when R410a is used as the primary cooling medium202, the vapor pipes222may be sized to allow a velocity of the primary cooling medium202in the vapor phase206to be preferably less than 1,000 fpm and more preferably less than 600 fpm. These header design features are not limited to microchannel coils but may also apply to other evaporators and condensers including finned tube coils.

As discussed above, the cooling system110of this invention may include a plurality of airstream cooling assembly loops200. For example, the cooling system110shown inFIGS.1and2has four airstream cooling assembly loops200. In the following discussion of multiple loops, the same reference numerals are used as discussed above with reference toFIGS.3and4, and letters are appended to the reference numerals to designate the different loops. For example, the letter “a” is appended to components of a first airstream cooling assembly loop200a, the letter “b” is appended to components of a second airstream cooling assembly loop200b, etc.

One pair of evaporators212a,212bis arranged in parallel with another pair of evaporators212c,212drelative to the return air114airstream. The evaporators212a,212b,212c,212dwithin each pair are arranged in series. In the first pair, return air114is directed across a first evaporator212aof the pair before being directed across a second evaporator212bof the pair. Likewise, in the second pair, return air114is directed across a first evaporator212cof the pair before being directed across a second evaporator212dof the pair. The corresponding passive condensers214a,214b,214c,214dare similarly arranged in pairs (first pair214a,214band second pair214c,214d) with the first pair being in parallel with the second pair and each condenser within the pair being arranged in series. In the first pair, scavenger air118is directed across the first condenser214aof the pair before being directed across the second condenser214bof the pair and, in the second pair, across the first condenser214cof the pair before being directed across the second condenser212dof the pair.

In one configuration of the cooling system110shown inFIGS.1and2, the cooling system110may be enclosed with a footprint of 32 feet long (FIG.1) and 10 feet, 2 inches wide (FIG.2) with a total height (excluding the scavenger fans120) of 12 feet, 8 inches. In this example, there is 88 ft2of coil area (evaporator212) available for the cooling process. At a nominal 500 fpm through the evaporators212, a design flow of 44,000 scfm and 348 kW is possible, resulting in a perimeter watt capacity of 107 kW/m. The coil length of the evaporator212can easily be extended to further increase the airflow and thus the capacity of the cooling system110without increasing its width, making even greater perimeter watt capacities possible. The cooling system110may be divided into two sections, an interior air handler132and a condensing unit134. As shown inFIGS.1and2, the evaporators212a,212b,212c,212dand the supply air fans116are located in the interior air handler132. The remaining components of the airstream cooling assembly200, including the secondary system230and scavenger air fans120, are located in the condensing unit134. InFIG.1, the condensing unit134is shown adjacent to the interior air handler132, but it can be at any suitable location including, for example, on top of the roof of data center100(e.g., the building housing the racks102).

Any number of suitable configurations for multiple airstream cooling assembly loops200may be used. For example,FIG.6is a schematic of another arrangement for the cooling system110having a plurality of airstream cooling assembly loops200. In the configuration shown inFIG.6, the cooling system110has four airstream cooling assembly loops200. The evaporators212e,212f,212g,212hof each loop200are arranged in series with respect to the return air114, but as discussed above the evaporators212e,212f,212g,212hmay also be arranged in parallel. The return air is first directed over the evaporator212eof the first loop before being subsequently directed over the evaporator212fof the second loop, the evaporator212gof the third loop, and the evaporator212hof the fourth loop, in that order. In the configuration shown inFIG.6, all four passive condensers214e,214f,214g,214hare arranged in parallel with respect to the scavenger air118, but as discussed above the passive condensers214e,214f,214g,214hmay also be arranged in series. Each condenser234e,234f,234g,234hfor the secondary cooling system230(which in this embodiment is a direct expansion cooling system) is arranged in series, relative to the scavenger air118, with the passive condenser214e,214f,214g,214hof the corresponding loop.

In general, the internal temperature of each airstream cooling assembly loop200will be isothermal, but each of the four airstream cooling assembly loops200will operate at a different temperature and pressure. The temperature of the primary cooling medium202in the first loop will be the warmest as air entering the evaporator212eof the first loop will be the warmest (the initial temperature of the return air114). The air entering the evaporators212f,212g,212hbecomes subsequently cooler than that in the previous loop because of the cooling resulting from the previous loop. When the ambient air temperature is lower than the temperature of each of the airstream cooling assembly loops200, the energy in the primary cooling medium can be transferred from the return air114to the scavenger air118in the passive mode with all four loops operating in passive mode as shown inFIG.6.

Each loop can be selectively operated in either the passive or active mode.FIG.7shows the cooling system110operating with the fourth loop in active mode and the other three loops operating in passive mode, andFIG.8shows all four loops operating in active mode. A controller240may be used to operate the cooling system110. In this embodiment, the controller240is a microprocessor-based controller that includes a processor242for performing various functions discussed further below and a memory244for storing various data. The controller240may also be referred to as a CPU. In one embodiment, control of the cooling system110may be implemented by way of a series of instructions stored in the memory244and executed by the processor242.

The controller240is communicatively coupled to a temperature sensor (“TS”)122. In this embodiment, the temperature sensor122is used to monitor the temperature of the supply air112, allowing the temperature sensor122to transmit (and the controller240to receive) the temperature of the supply air112. Loop sensors250may also be used to measure various parameters of each airstream cooling assembly loop200. For example, the loop sensors250may measure the temperature and pressure of the primary cooling medium202in each loop using a temperature sensor (“TS”)252and a pressure sensor (“PS”)254, respectively. Preferably the temperature and pressure sensors252,254are located in the liquid refrigerant line224to monitor the temperature and pressure of the liquid204phase of the primary cooling medium202.

The controller240may also be communicatively coupled to other components of the cooling system110and used to control those components as well. For example, the supply air fans116and the scavenger air fans120may be communicatively coupled to the controller240, and thus the controller240may be used to direct the return air114and scavenger air118over the evaporators212e,212f,212g,212hand condensers214e,214f,214g,214h, respectively, and increase or decrease the airflow. The controller240may also be communicatively coupled to the secondary cooling system230e,230f,230g,230hof each loop and used to turn on or off (activate or deactivate) the secondary cooling system230e,230f,230g,230h.

FIG.9is a flow chart showing one example of how the cooling system110shown inFIGS.6and7may be controlled. In step S405, the controller240directs the return air114over the evaporators212e,212f,212g,212h. The supply air temperature sensor122is used to measure the temperature of the supply air112, and the controller240receives the temperature of the supply air112in step S410. The controller240then compares the measured temperature of the supply air112with a set point in step S415. The set point may be provided to the controller240using any suitable method or device. For example, the controller240may be communicatively coupled to a user interface through which a user can provide a desired temperature of the supply air112, and the controller240can receive the desired temperature of the supply air112to use as the set point. If the temperature of the supply air112is equal to the set point (or within a suitable operating range of the set point), the control system240returns to step S405to continue monitoring the temperature of the supply air112.

If the temperature of the supply air112is too cold (below the temperature of the set point or operating range), the controller240checks, in step S420, to see if any of the loops200are operating in the active mode. The controller240may store, for example, the mode of a loop in the memory244when the controller activates or deactivates the secondary cooling system230for that loop. The controller240may then query the memory244to determine the mode of any loop. The controller240may store in the memory244other suitable parameters, such as the flow rate of the scavenger air118(e.g., speed and number of scavenger fans120running) for example, the controller240may likewise check and alter these parameters in a similar manner. If the controller240determines (in step S420) that no loops200are in the active mode, the controller240reduces the airflow of the scavenger air118in step S425before returning to step S405to continue monitoring the temperature of the supply air112. If any changes to the cooling system110are made in step S425(or any other step discussed herein), the controller240may delay monitoring the temperature of the supply air112to allow the change to impact the temperature of supply air112.

If the controller240determines, in step S420, that at least one loop200is in the active mode, the controller240deactivates the secondary cooling system230in one of the loops200, in step S430. As shown inFIG.7for example, the fourth loop is operating in the active mode. If the temperature of the supply air is too cold in this configuration, the controller240would deactivate the secondary cooling system230hof the fourth loop, returning the fourth loop to the passive mode, as shown inFIG.6. Preferably, the controller240will deactivate the secondary cooling system230of the loop operating in the active mode which has its evaporator212the farthest upstream relative to the return air114. The controller240then returns to step S405to continue monitoring the temperature of supply air112.

If the temperature of the supply air112is too hot (above the temperature of the set point or operating range), the controller240first checks, in step S435, to see if the airflow of the scavenger air118can be increased. If the airflow of the scavenger air118can be increased (the airflow the scavenger air118is not at its maximum), the controller240increases the airflow of the scavenger air118in step S440before returning to step S405. The controller240may increase the airflow of the scavenger air118by any suitable means including, for example, by increasing the speed of the scavenger air fans120. If the airflow of the scavenger air118cannot be increased (the airflow of the scavenger air118is at its maximum), the controller240checks, in step S445, to see if all the loops200are in the active mode. If all of the loops200are in the active mode, as shown inFIG.8, the cooling system110is operating at its maximum cooling capacity and the controller240returns to step S405. If at least one loop is in the passive mode, the controller240will activate a secondary cooling system230for one of the loops200in step S450. For example, if all of the loops are operating in the passive mode as shown inFIG.6, the controller240will activate the secondary cooling system230of one of the loops200, such as the secondary cooling system230hof the fourth loop. Preferably, the controller240will activate the secondary cooling system230of the loop operating in the passive mode which has its evaporator212the farthest downstream relative to the return air114. The controller240then returns to step S405to continue monitoring the temperature of supply air112.

For data center cooling systems, it is often desirable to have efficiencies of 65% or greater in an economizer mode (passive mode in this embodiment). In the passive mode, the refrigerant is at virtually the same pressure at all locations within the primary coolant loop210, and the internal temperature is isothermal. Based on energy balance requirements, if the passive condenser214and the evaporator212heat transfer constraints are the same (equal air flows over the outer surface of the condenser214and evaporator212and surface characteristics of the outer surface of the condenser214and evaporator212), the refrigerant would exist at a temperature equal to the average of the evaporator212and the passive condenser214inlets, and, in a non-ideal world, the net efficiency of a single loop200would be less than 50%. A heat exchange efficiency, when measured on the evaporator size of greater than 50%, is achievable, however, with unbalanced airflows over the outer surface of the condenser214and evaporator212.

Using multiple loops200with air counterflow to the flow in the primary coolant loop210, the efficiencies of each loop will have additive effect and efficiencies greater than single loop efficiencies can be achieved. For example, if two loops200, each having an efficiency of 50%, are used with the scavenger air118flowing in series through the first loop and then the second loop and the return air114flowing in the opposite direction (through the second loop and then the first loop) an efficiency greater than 70% can be achieved. If, however, the efficiency for a single loop200drops to 39%, three loops200, rather than two, can be positioned in counterflow to achieve a net efficiency greater than 70%. The calculations above used scavenger air118having a temperature of 70° F. delivered at 10,000 cfm and return air114having a temperature of 100° F. delivered at 5,000 cfm.

The following examples (Cases 1 through 6) were constructed to evaluate the efficiency of a single loop200. The results of these evaluations are presented in Table 1 below. The following cases used an unbalanced airflow where the scavenger fan120was selected to provide at least a 2:1 airflow ratio of scavenger air118to return air114based on a nominal 500 fpm face velocity of the return air114over the evaporator212. In the following experimental cases, however, a flow ratio closer to 2.2:1 was achieved with the total airflow across the evaporator212being 5,000 scfm and the total airflow across the passive condenser214being 11,000 scfm. The face velocity across the passive condenser214was 500 fpm.

The first case (Case 1) used finned tube (“FT”) coils for both the evaporator212and passive condenser214. The coil for the evaporator212was a flooded, two-row, one-pass coil, and the coil for the passive condenser214was a three-row, one-pass coil. Both coils used half-inch tubes in a typical tube arrangement and had 10 fins per inch. Each coil was 5 ft. long. Both the evaporator212and the passive condenser214were mounted at a 15 degree angle relative to horizontal to facilitate vapor expulsion and condensate flow via gravity. The passive condenser214was mounted with its lower end 2 feet above the upper discharge of the evaporator212. Vapor and liquid lines between the coils were oversized, using a 1⅛ inch pipe for the liquid refrigerant line224and a 2⅛ pipe for the vapor pipe222, so as not to inhibit refrigerant flow and affect the resultant performance. R410a was used as the refrigerant.

The second case (Case 2) was the same as the first case, but a flooded microchannel coil (MC) was used as the evaporator212instead of a finned tube coil. Using a microchannel coil significantly reduced the refrigerant charge needed because the internal volume of the microchannel coil was greatly reduced (by over 47%) compared to the half-inch tube coil. Each microchannel extrusion330had a width of 38 mm with 28 microchannels334,336. The width of each of the26internal microchannels334was 0.92 mm and the outer two microchannels336(seeFIG.5B) was rounded with a 0.55 mm radius and had an overall width of 0.94 mm. The microchannel extrusion330had an overall height of 1.8 mm with an external wall thickness t of 0.35 mm. The internal wall thickness separating the microchannels334was 0.40 mm. Sixty-seven microchannel extrusions330were used and each had a length of 1.57 m. A single liquid connection312with an outside diameter of 22.2 mm was used and a single vapor connection322with an outside diameter of 25 mm was used.

The third case (Case 3) was the same as the second case, but a microchannel coil (MC) was used as the passive condenser214instead of a finned tube coil. The microchannel coil for the passive condenser214was designed similarly to the microchannel coil for the evaporator212(described above in Case 2), but the passive condenser214used 100 microchannel extrusions330, each having a length of 1.57 m.

The fourth case (Case 4) was the same as the third case, but three vapor pipes222and vapor connections322were used instead of one (MC Mod). The configuration of the loops200in Cases 3 and 4 was also evaluated with a higher temperature differential between the evaporator212and the passive condenser214(Cases 5 and 6, respectively). The temperature differential between the evaporator212and passive condenser214in Cases 5 and 6 was increased by increasing the temperature of the return air114by 20° F. to 25° F., relative to Cases 3 and 4. Table 1 below presents the results of each of the cases and in the table “Evap” refers to the evaporator212and “Cond” refers to the passive condenser214.

As can be seen in Table 1 above by comparing Case 1 to Case 2, using a microchannel coil instead of a finned tube coil in the evaporator212made the most significant difference to performance, increasing heat exchange efficiency to from 34% to 57%. Changing the passive condenser214to a microchannel coil had little effect on the performance results at normal conditions (compare Case 2 with Case 3). Modifying the evaporator212and passive condenser214to have additional vapor connections322resulted in a 3% efficiency gain at normal conditions (compare Case 3 with Case 4). When the temperature difference between the evaporator212and passive condenser214was increased, the efficiency dropped (compare Cases 3 and 4 with Cases 5 and 6, respectively), but having multiple vapor connections322resulted in a smaller efficiency drop and the total power transfer increased greatly to nearly 50 kW.

Using the properties of R410a and a known heat transfer rate, a mass flow of primary cooling medium202may be calculated based on the specific heats of the vapor and liquid. In the case of the microchannel evaporator212, the limit to heat flux was in the range of 20 kW. Using an enthalpy difference between liquid and vapor for R410a, a mass flow of 387 kg/hr and flow rate of 5.88 m3/hr are calculated. Applying the inside diameter of a single ⅞ inch tube, the velocity of the gas is 4.2 m/s. Adding two additional vapor connections to the coil increases the capacity 50 kW and results in a velocity of 3.1 m/s. So for practical purposes, when using R410a, piping connections are preferably sized for a maximum velocity under about 4 m/s. The vapor flow rate inside the microchannel extrusions330is 2.1 m/s for the heat exchange rate of 50 kW. The sizing will vary for other refrigerants based on their density and viscosity but can be determined experimentally.

A plate heat exchanger was used as the active condenser216in each of Cases 1 through 6. The active condenser216was arranged in parallel with the passive condenser214, and chilled water was used as the secondary cooling medium208. In the active mode, the efficiency data and maximum power data closely mimicked the air to air data, confirming the superiority of the microchannel evaporator212to the finned, tube evaporator212and the increase in total capacity after the addition of the extra header connections to the microchannel for the vapor transport.

Another cooling system110is shown inFIGS.10and11. Features and components of the cooling system110shown inFIGS.10and11are similar to those discussed above inFIGS.1-9. The same reference numerals are used inFIGS.10and11to describe the same and similar components as those discussed above, and a detailed description of these components is omitted from the following discussion.FIG.10shows the interior air handler132andFIG.11shows the condensing unit134of the cooling system110of this embodiment.

As with the cooling system110shown inFIG.1, the cooling system110shown inFIGS.10and11includes a plurality of evaporators212, which in this embodiment is four evaporators212i,212j, two first evaporators212iand two second evaporators212j. The first evaporators212iare arranged in parallel relative to the return air114with one another and are connected to both a first common vapor pipe262and first common liquid refrigerant line266. Likewise, the second evaporators212jare arranged in parallel relative to the return air114with one another and are connected to both a second common vapor pipe264and a second common liquid refrigerant line268. Each of the first common liquid refrigerant line266and second common liquid refrigerant line268are similar to the vapor pipe222and the liquid refrigerant line224, respectively, as discussed above. Each first evaporator212iis arranged in series with one of the second evaporators212jrelative to the return air114. Return air114is directed across the first evaporator212ibefore being directed across the second evaporator212j.

The condensing unit134shown inFIG.11includes four circuits, a first circuit (Circuit1), a second circuit (Circuit2), a third circuit (Circuit3), and a fourth circuit (Circuit4). In the following discussion of each of these circuits, the same reference numerals are used as discussed above with reference toFIGS.3and4, and reference characters are appended to the reference numerals to designate the different circuits. Reference character “c1” is appended to components of the first circuit. Reference character “c2” is appended to components of the second circuit. Reference character “c3” is appended to components of the third circuit. Reference character “c4” is appended to components of the fourth circuit. Where the discussion of the component is common to any of the circuits, however, the reference character referring to a particular circuit is omitted. Although the condensing unit134is described herein as having four circuits, any suitable number of circuits may be used. In addition, although described with specific components in each of these four circuits (for example passive condensers214and active condensers216), various arrangements of these components are contemplated to be within the scope of the invention.

Additionally, in each of the first, second, and third circuits, two passive condensers214are associated with the first evaporators212i(part of a first thermosiphon loop200i) and two passive condensers214jare associated with the second evaporators212j(part of a second thermosiphon loop200j). These condensers214will also have either “i” or “j” appended to the reference numerals to designate in which loop the passive condenser214is located. For example,214ic1is used to designate one of two passive condensers214in the first circuit that are part of the first thermosiphon loop200i.

As shown inFIG.11, the first evaporators212iare connected to six passive condensers214iin parallel. Two of the six passive condensers214iare located in each of the first, second, and third circuits. The first thermosiphon loop200iincludes one active condenser216c4in parallel with the six passive condensers214i. The active condenser234c4of the first thermosiphon loop200iis located in the fourth circuit. In this embodiment, the fourth circuit includes the active condenser216c4and its corresponding secondary cooling system230c4, but does not include any passive condensers214. Scavenger air fans120c4are configured to direct scavenger air118over the outer surface of the condensers234c4of the secondary cooling system230c4.

The second evaporators212jare also connected to six passive condensers214jin parallel. Two of the six passive condensers214jare located in each of the first, second, and third circuits. The second thermosiphon loop200jincludes three active condensers216j. One of each of the three active condensers216jc1,216jc2,216jc3is located in each of the first, second, and third circuits. Each of the four circuits thus includes a secondary cooling system230. In this embodiment, each of the secondary cooling systems230c1,230c2,230c3,230c4includes two condensers234connected in parallel to each other. In other embodiments, a pre-cooler124(seeFIG.12) may be used to cool the scavenger air118before it is passed through the passive condensers214jof the second thermosiphon loop200j. In such a case, it may be possible to omit the active condensers216jc1,216jc2,216jc3in the second thermosiphon loop200j.

The arrangement of each of the first, second, and third circuits are similar to each other. The following description of the first circuit applies equally to the second and third circuits. The condensers214ic1,214jc1,234c1in the first circuit are arranged in two sets, a first condenser set272c1and a second condenser set274c1. The first condenser set272c1and the second condenser set274c1are arranged in parallel with each other relative to the air flow of the scavenger air118. Each of the first condenser set272c1and the second condenser set274c1contain one of each of the passive condenser214jc1of the second thermosiphon loop200j, the passive condenser214ic1of the first thermosiphon loop200i, and the condenser234c1of the secondary cooling system230. The condensers214jc1,214ic1,234c1are arranged in series relative to the air flow of the scavenger air118. The scavenger air118is drawn by scavenger air fans120c1of the first circuit through each of the condensers as follows. The scavenger air118is ambient air drawn from the outdoor environment surrounding the condensing unit134and is first passed through the passive condenser214jc1of the second thermosiphon loop200j. Next, the scavenger air118is passed through the passive condenser214ic1of the first thermosiphon loop200i. Then, the scavenger air118passes through the condenser234c1of the secondary cooling system230before being exhausted to the outside by the scavenger fans120. Each of the scavenger air fans120may be independently variable or at least variable between different circuits.

This arrangement of condensers214jc1,214ic1,234c1in the first circuit allows for a counter flow design. The primary cooling medium202in the second thermosiphon loop200jis cooler than the primary cooling medium202in the first thermosiphon loop200i. Thus, the coldest scavenger air118passes through the coldest condenser214jc1first, and then after being heated by the passive condenser214jc1of the second thermosiphon loop200j, scavenger air118passes through the warmer passive condenser214ic1of the first thermosiphon loop200j.

The cooling system110shown inFIGS.10and11, like the cooling systems110discussed above, does not use valves in the first thermosiphon loop200iand second thermosiphon loop200jto switch between the active and passive modes. Instead, by activating the secondary cooling system230to cool the active condenser216, the vapor206of the primary cooling medium202naturally travels to the colder active condenser216to condense, and the airstream cooling assembly100, through thermodynamic forces, moves from passive mode to active mode.

The cooling system110shown inFIGS.10and11may be controlled similarly to the cooling systems110described above, such as using the process shown and described with reference toFIG.9. The temperature of the supply air112may be controlled to a set point (see step S415). First, if the temperature of the supply air112is above the set point, the controller240that is used to control the cooling system110will increase the flow rate of the scavenger air118, such as by increasing the fan speed of the scavenger fans120(see step S440). If the temperature of the supply air112is below the set point, the controller240will decrease the flow rate of the scavenger air118(e.g., decrease the fan speed of the scavenger fans120) (see step S425).

If the temperature of the supply air112is above the set point and the flow rate of the scavenger air118is at its maximum, the controller240will then energize a secondary cooling system230(see step S450). In the cooling system110shown inFIGS.10and11, the controller240may stage on the secondary cooling system230by circuits as necessary to maintain the temperature of the supply air112at the set point. Although the secondary cooling system230of the circuits may be staged on in different sequences, one approach is to stage on the secondary cooling systems230successively, as needed, in the order of the first circuit, then the second circuit, then the third circuit, and then the fourth circuit. Thus, in this embodiment, the secondary cooling systems230of the second thermosiphon loop200jmay be staged on before the secondary cooling system230c4of the first thermosiphon loop200i. If the temperature of the supply air112is below the set point, the controller240may then deactivate a secondary cooling system230(see step S430) in, for example, the opposite order as the circuits are staged on.

If the temperature of the supply air112is below the set point and all secondary cooling systems230(active cooling modes) are off with the fan speed of the scavenger fans120at a minimum, the controller240may then stage off scavenger fans120as necessary to maintain the temperature of the supply air112at the set point. In a case where all but one of the scavenger fans120is off, the controller240may operate only one of the first thermosiphon loop200iand the second thermosiphon loop200j. This can be achieved by closing the flow control valves276(discussed further below) for the loop that has been deactivated, which, for example, may be the second thermosiphon loop200j.

Even when some or all of the active condensers216are being operated, and thus the system is operating in the active mode, the inventors have unexpectedly found that there is some economization (cooling of the primary cooling medium202from the passive condensers214) even in the active mode, as demonstrated by the test results in Table 2, below. Table 2 shows the results of a test with a nominal heat load of 350 kW over the evaporators212. At least one circuit of the first thermosiphon loop200iand second thermosiphon loop200jis operating in the active mode for the conditions in the table below. The test was run for various different ambient air (scavenger air118) temperatures and the heat rejection from both the active condensers216and passive condensers214was measured. The active modes of the circuits were staged on and off based on the ambient air temperature, but at the highest ambient air temperature all of the secondary cooling systems230were operating.

Heat⁢⁢Rejection⁢⁢(kW)=V⁡(TI-TO)5.9605×fa⁡(TO+459.67)(1)
In equation (1), V is the volume of the air (such as the supply air112) in actual cubic feet per minute across a condenser or an evaporator, Ti is the temperature, in Fahrenheit, of the air going into the condenser or the evaporator (such as the return air114), To is the temperature, in Fahrenheit, leaving the condenser or the evaporator (such as the supply air112), and fais a factor based on the altitude. The altitude factor (fa) may be calculated using equation (2) below, where A is the altitude in feet.
fa=(1−(6.8754×106)A)−5.2559(2)

Although valves are not used to switch between modes, valves can be used to help regulate the cooling system110. In this embodiment, a flow control valve276is placed at the entrance to each evaporator212in the first common liquid refrigerant line266and the second common liquid refrigerant line268. Any suitable valve may be used as the flow control valve276, but in this embodiment the flow control valve276is a globe valve operated by a stepper motor. Here, the flow control valve276allows for continuous adjustment of flow past the plug or disk of the flow control valve276and the plug or disk is operated by the stepper motor. The flow control valve276valve includes a closed position and a plurality of open positions each having a different opening area. The plurality of open positions thus allow a different amount of the primary cooling medium202in the liquid phase204to flow past the flow control valve276based on the opening area of that position.

The flow control valve276is used to precisely control the liquid level in the evaporator212(amount of primary cooling medium202in the liquid phase204) and maintain a desired temperature of the vapor206leaving the evaporator212. The inventors have found that using the flow control valve276allows the primary cooling medium202to efficiently circulate through the first thermosiphon loop200iand the second thermosiphon loop200jby natural circulation and gravity for a wide range of heat loads and ambient air conditions. The flow control valve276can be used to prevent too much liquid204from entering the evaporator212(e.g., flooding the evaporator212), which could inhibit vapor206flow out of the evaporator212. Further, the flow control valve276can be used to prevent too little liquid204from entering the evaporator212(e.g., starving the evaporator212), which could inhibit effective and efficient condensing in the condensers214,216. Such considerations, and thus the use of the flow control valve276, may be particularly relevant where the interior air handler132and the condensing unit134are separated, as greater distances require larger amounts of the primary cooling medium202, further exacerbating the issues discussed above such as flooding.

Various approaches may be used to set the position of the flow control valve276and thus the amount of liquid204flowing into the evaporator212. For example, the position of the flow control valve276may be based on heat absorption in the evaporator212, heat rejection of the return air114/supply air112, heat rejection in the condensers214,216, heat absorption by the scavenger air118, or superheat of the vapor206. As discussed above, the controller240is communicatively coupled to various sensors such as the loop sensors250(e.g., temperature sensors252and pressure sensors254) located in the first thermosiphon loop200iand the second thermosiphon loop200jor temperature sensors used to monitor the temperature of the supply air112, return air114, and scavenger air118. Using outputs from these sensors (inputs into the controller), the controller240can determine the appropriate position of the flow control valve276and drive the stepper motor of the flow control valve276as appropriate.

When controlling the flow control valve276based on heat absorption in the evaporator212, the heat absorption may be determined based on the temperature rise across the evaporator212. One way to determine temperature rise is to measure the temperature of the primary cooling medium202at the inlet of the evaporator212(or another suitable location at the bottom of the evaporator212) and the temperature of the primary cooling medium202at the outlet of the evaporator212(or another suitable location at the top of the evaporator212). Another way to determine temperature rise is to measure the temperature of the supply air112leaving the bottom third of the evaporator212(the third of the evaporator212proximate the inlet) and the temperature of the supply air112leaving the top third of the evaporator212(the third of the evaporator212proximate the outlet). The temperature rise can then be determined by taking the difference between the measured temperatures and comparing the difference to a set point. If the difference is lower than the set point, the controller240controls the stepper motor to move the flow control valve276in a closed direction to reduce the flow of liquid204into the evaporator212. If the temperature is higher than the set point, the controller240controls the stepper motor to move the flow control valve276in an open direction to increase the flow of liquid204into the evaporator212.

Another way to control the flow control valves276is to measure the heat rejected by the return air114/supply air112. When controlling the flow control valve276based on heat rejection of the return air114/supply air112, the temperature of the return air114may be measured before it reaches the evaporator212and then again after passing through the evaporator212using a temperature sensor. The controller240may then be used to calculate an amount of heat rejection in that thermosiphon loop (e.g., the first thermosiphon loop200ior the second thermosiphon loop200j) based on the measured temperatures and flow rate of the return air114/supply air112. The controller then sets the position of the flow control valves276in the thermosiphon loop200i,200jas a function of the amount of heat absorbed in that thermosiphon loop200i,200jbased on curves or a look-up table. Different curves or values could be used for different operating modes, such as a curve for when each circuit is operating in an active mode.

When controlling the flow control valve276based on heat rejection in the condensers214,216, the heat rejection may be determined based on the temperature drop across each of the condensers214,216. As with measuring temperature rise, as discussed above, one way to determine temperature drop is to measure the temperature of the primary cooling medium202at the inlet of a respective condenser214,216(or another suitable location at the top of the condenser214,216) and the temperature of the primary cooling medium202at the outlet of the condenser214,216(or another suitable location at the bottom of the condenser214,216). Another way to determine temperature rise is to measure the temperature of the scavenger air118leaving the top third of the condenser214,216(the third of the condenser214,216proximate the inlet) and the temperature of the scavenger air118leaving the bottom third of the condenser214,216(the third of the condenser214,216proximate the outlet). The temperature drop can then be determined by taking the difference between the measured temperatures. The temperature drop for each condenser214,216may be used to calculate a total amount of heat rejected by the condensers214,216in that thermosiphon loop (e.g., the first thermosiphon loop200ior the second thermosiphon loop200j) and the controller240then sets the position of the flow control valves276in the thermosiphon loop200i,200jas a function of the amount of heat rejection in that thermosiphon loop200i,200jbased on curves or a look-up table. As discussed above, different curves or values could be used for different operating modes, such as a curve for when each circuit is operating in an active mode. In such a case, the flow control valves276can be adjusted, as discussed above, to be more open with more heat rejection or more closed with less heat rejection. Alternatively, instead of using the controller240to calculate heat rejection, the temperature drop may be used directly as the basis of the curves or the look-up table.

Another way to determine the amount of heat rejected by the condensing unit134is to measure the heat absorbed by the scavenger air118. When controlling the flow control valve276based on heat absorption by the scavenger air118, the temperature of the scavenger air118may be measured before it reaches the condensers214,216,234and after passing through the condensers214,216,234using a temperature sensor. The controller240may then be used to calculate an amount of absorption in that thermosiphon loop (e.g., the first thermosiphon loop200ior the second thermosiphon loop200j) based on the measured temperatures and flow rate of the scavenger air118, and the controller then sets the position of the flow control valves276in the thermosiphon loop200i,200jas a function of the amount of heat absorbed in that thermosiphon loop200i,200jbased on curves or a look-up table, in a manner similar to that discussed above for controlling the flow control valve276based on heat rejection in the condensers214,216.

When controlling the flow control valve276based on a predetermined value of superheated vapor206(a set point), the flow control valve276can be opened to allow more condensed liquid204into the evaporator212, thereby reducing the superheat temperature of the vapor206, if the superheat temperature of the vapor206rises above the set point. Similarly, if the superheat temperature of the vapor206is reduced to below the set point, the flow control valve276can be closed, reducing the flow of condensed liquid204into the evaporator212, thereby increasing the superheat temperature of the vapor206. Loop sensors250, such as a temperature sensor252and a pressure sensor254, may be located at the outlet of the evaporator212. Such sensors250also may be suitably located in the evaporator212itself or in the vapor lines leading to the condensers214,216. The controller240then calculates the superheat based on the temperature and pressure measurements and compares it to the set point (e.g., a predetermined (desired) level). The controller240then adjusts the flow control valve276, as discussed above. In any of the control methods, the adjustment may not be made as soon as the measured or calculated values cross the set point; rather the controller240may adjust the flow control valve276when an upper threshold temperature (or value) and a lower threshold temperature (or value) that are above and below the set point, respectively, are crossed. While the invention is described with controls based on set points herein, a person of ordinary skill in the art would understand that these operating bands are included by such a description.

As with any of the cooling systems110described herein, the speed of the supply air fans116that are driving the airflow (return air114) across the evaporators212may be varied. The supply air fans116may be downstream of the evaporators212(in the direction of travel of the return air114/supply air112) as shown inFIG.10. Alternatively, the supply air fans116may be upstream of the evaporators212, as shown inFIG.1, and the supply air fans116may also be referred to as return air fans. In one instance, the air flow of the return air114/supply air112may be driven by the requirements of the data center100. The data center100may have a controller (not shown), referred to herein as the building management system (BMS), that is separate from the controller240for the cooling system110. The BMS determines what the speed of the supply air fans116should be based on the needs in the data center100. Such needs may vary from data center to data center. The BMS may send a supply fan speed signal to the controller240for the cooling system110and the controller240may modulate the supply air fans116(e.g., speed and number of supply air fans116running) to match the signal given by the BMS.

Another way to control the supply air fans116in the cooling system110, such as if the BMS signal is absent, is to control the supply air fans116to a return air temperature set point for the temperature of the return air114. The controller240can modulate the flow rate of the return air114/supply air112by adjusting the speed of the supply air fans116(or number of supply air fans116operating) to maintain the temperature of the return air114to the return air temperature set point. If the temperature of the return air114is above the return air temperature set point, the speed (or number) of the supply air fans116is increased. If the temperature of the return air114is below the return air temperature set point, the speed (or number) of the supply air fans116is decreased.

The fluid being cooled by the cooling systems110discussed herein may be referred to as a process fluid142. In the embodiments discussed above, the process fluid142cooled by the cooling systems110is air. In the previous discussions, air (process fluid142) is directed over racks102containing electronics and heated before being directed over an evaporator212(return air114) to be cooled. The cooling systems110described herein are not limited to cooling air, however, and may be used to cool any suitable fluid. The process fluid142may include, for example, liquids such as water, water and glycol mixtures, and a non-conductive fluid (dielectric).

In the embodiments discussed above, where the process fluid142is air, the evaporator212was suitably a microchannel coil or finned tube coils. Where the process fluid142is a liquid instead of a vapor (gas), other suitable evaporators212may be used, including, for example, a plate heat exchanger, a coaxial heat exchanger, or a shell and tube heat exchanger.FIG.12shows a cooling system110in which the evaporators212are suitably designed to cool a process fluid142that is a liquid. The process fluid142is circulated in a process fluid loop140by a pump144. This embodiment includes a first evaporator212kand a second evaporator212llocated in series in the process fluid loop140. Process fluid142is heated by a heat load such as servers103in a rack102(seeFIGS.13A and13B). The process fluid142is then cooled by the first evaporator212kand the second evaporator212lbefore being returned to cool the servers103.

In this embodiment, the second evaporator212lis fluidly connected to both a passive condenser214(passive condenser214l) and an active condenser216(active condenser216l) and operates like the loop200described above inFIGS.3and4. The first evaporator212kis fluidly connected to a passive condenser214(passive condenser214k), which is located downstream of the passive condenser214lrelative to the air flow direction of the scavenger air118between the passive condenser214land the condenser234of the secondary cooling system230. As with the embodiments discussed above, the first evaporator212kmay also be fluidly connected to both the passive condenser214kand an active condenser216(active condenser216k) to operate like the loop200described above inFIGS.3and4. The condensing unit134of this embodiment may also include an adiabatic pre-cooler124to pre-cool the scavenger air118before it passes through any of the condensers214,234. Any suitable adiabatic pre-cooler124may be used including, for example, Munters FA6™ Evaporative Humidifier/Cooler manufactured by Munters Corporation of Buena Vista, Virginia, USA.

As discussed above, the process fluid142may be heated by a heat load such as servers103located in a rack102of the data center100. The cooling system110shown inFIG.12may be suitable for use with immersion cooling systems for servers103.FIGS.13A and13Bshow examples of a rack102used in an immersion cooling system. In FIG.13A, the servers103are submerged in a dielectric146. The servers103heat the dielectric146, but the dielectric146remains a liquid (single phase). The heated dielectric146is circulated as the process fluid142through the process fluid loop140to be cooled and returned to the rack102to further cool the servers103.

InFIG.13B, the servers103are also submerged in the dielectric146. In this case, the dielectric146cools the servers103by two-phase cooling. The servers103heat the dielectric146, and the dielectric146changes phase into a vapor (gas). The dielectric vapor rises to the top of the rack102. The top of the rack102includes a coil148. An appropriate process fluid142passes through the coil148and condenses the dielectric146. In another embodiment, the dielectric146in the vapor phase may be directly cooled by the second evaporator212land the first evaporator212kas the process fluid142instead of using another fluid.

In the embodiments described above the servers103are physically separated from the first evaporator212kand second evaporator212land the process fluid loop140is used to transport heat from the servers103or other information technology (“IT”) equipment. The inventions described herein are not so limited, however, and the evaporators212may be any liquid to refrigerant heat exchanger, where a circulating liquid (dielectric fluid, water, or other fluid) transports heat from the IT equipment to the refrigerant that is integral to the two-phase thermosiphon loop. Such other suitable evaporators212include, for example, a cold plate integrated into the servers103or IT component to directly absorb heat from the component and/or chips therein or a plurality of tubular surfaces directly integrated into a submersion cooling system.

A second preferred embodiment of an airstream cooling assembly loop500is shown inFIGS.14and15. In this embodiment, the evaporator512and the passive condenser514are incorporated in an integral heat exchanger510, which operates as a heat pipe. The evaporator512is a lower portion of the integral heat exchanger510and may also be referred to as an evaporator portion512. Likewise, the passive condenser514is an upper portion of the integral heat exchanger510and may also be referred to as a condenser portion. As with the first embodiment, any suitable heat exchanger may be used, including finned tube coils or microchannel coils. In this embodiment, the integral heat exchanger510is shown as a finned tube coil with tubes516connecting two fixed headers, a top header522and a bottom header524. As will be described below, gravity also plays a role in the cooling process and as a result, the tubes516are oriented preferably upright and more preferably, vertically.

The airstream cooling assembly500operating in passive mode is shown inFIG.14. The hot return air114is directed over the evaporator portion512of the integral heat exchanger510. The primary cooling medium202, which is contained within the tubes516, changes from the liquid phase204to the gas phase206, drawing heat from the return air114and, as a result, cooling the return air114. As the primary cooling medium202evaporates, the vapor206rises in the tube516to the condenser portion514of the heat exchanger. In the passive mode, scavenger air118is directed over the condenser portion514. Heat is rejected from the primary cooling medium202to the scavenger air118, condensing the primary cooling medium from the gas phase206to the liquid phase204. The liquid204of the primary cooling medium202then wicks down the sides of the tube516, assisted by gravity, back to the evaporator portion512.

The airstream cooling assembly500operating in active mode is shown inFIG.15. As with the airstream cooling assembly200of the first embodiment, the evaporator512of the airstream cooling assembly500is also connected to a second condenser, an active heat exchanger condenser216. A vapor pipe526connects the top header522of the integral heat exchanger510with the active condenser216. The vapor206of the primary cooling medium202travels through the vapor pipe526to the active condenser216. As in the first embodiment, heat is rejected from the primary cooling medium202to the secondary cooling medium208of the secondary cooling system230at the active condenser, causing the primary cooling medium202to change phase from vapor206to a liquid204. The condensed liquid204then travels, with the assistance of gravity, through a liquid refrigerant line528to the bottom header524, establishing a recirculated refrigerant flow.

Also like the first embodiment, the airstream cooling assembly500of this embodiment operates without the need of pumps, oils, compressors, or even valves to switch between modes. Instead, by activating the secondary cooling system230to cool the active condenser216, the vapor206of the primary cooling medium202naturally travels to the colder active condenser216to condense, and the airstream cooling assembly500automatically switches from passive mode to active mode. In addition, a controller240, which may be communicatively coupled to temperature sensors122,252, may be used to control the airstream cooling assembly500of this embodiment.

As discussed above, instead of a finned tube coil being used for the integral heat exchanger510a microchannel coil may be used. However, the total heat flux available may be limited in a passive mode that depends on heat pipe operation alone as the condensed liquid and evaporated gas flow counter to each other in the small channels of the microchannel extrusion.FIGS.16and17show another configuration of the second embodiment that includes a second passive condenser530(third condenser). In this embodiment, the third condenser530is a microchannel coil positioned in series, relative to the scavenger air118, with the condenser portion514of the integral heat exchanger510. Preferably, the third condenser530is positioned on the upstream side of the condenser portion514of the integral heat exchanger. The third condenser530has a vapor header532and a liquid header534. The vapor header532of the third condenser is connected to the top header522of the integral heat exchanger510by the vapor pipe526.

In the passive mode (shown inFIG.16), the vapor206of the primary cooling medium202will flow through the vapor pipe526and into the third heat exchanger, where most of the condensation of the primary cooling medium202will occur. As with the passive condenser214of the first embodiment, scavenger air118is driven across an outer surface of the third condenser530, and the heat of the primary cooling medium202contained in the third condenser530is released to the scavenger air118, condensing the vapor206to a liquid204. The liquid204of the primary cooling medium202then travels, with the assistance of gravity, through a liquid refrigerant line528to the bottom header524as a recirculated refrigerant flow.

In the active mode (shown inFIG.17), the vapor206naturally flows to the colder active condenser216to be condensed, as discussed above, and there is minimal, if any, flow of the primary cooling medium202through the third condenser530in the active mode. As with the first embodiment, it may be beneficial to include vapor traps (not shown) and/or check valves220in the liquid refrigerant lines528to avoid reverse flow of the primary cooling medium202through the condenser216,530that is not currently operating in a given mode.

The second passive condenser530may also be used in a configuration without the active condenser216, as shown inFIG.18. The configuration only has a passive mode, but the evaporator512is still connected to two condensers, the condensing portion514of the integral heat exchanger510and the second passive condenser530.

As with the airstream cooling assembly200of the first embodiment, the cooling system110may include a plurality of airstream cooling assemblies500of the second embodiment. For example, a cooling system110may include two loops500a,500bas shown inFIG.19. The first loop500ais similar to the configuration shown inFIG.18without an active condenser216, and the second loop500bis similar to the configuration shown inFIGS.14and15, but with a microchannel coil for the integral heat exchanger510b. In this configuration, the two evaporators512a,512bare arranged in series with respect to the return air114. The integral heat exchanger510bof the second loop500b(having the active condenser216) is located upstream of the integral heat exchanger510aof the first loop500a.

Another cooling system110is shown inFIGS.20and21having two airstream cooling assembly loops600, a first airstream cooling assembly loop600aand a second airstream cooling assembly loop600b, although any number of loops may be used, including a single loop. As with the discussion above, a letter is appended to a reference numeral to designate the loop in which the component is located. In this cooling system110, the return air114is directed across two cooling coils612a,612bthat are arranged in parallel with respect to the return air114, although the cooling coils612a,612bmay also be arranged in parallel. As the return air114flows across the cooling coil612, heat is transferred from the return air to a primary cooling medium602contained within a primary coolant loop610, heating the primary cooling medium602. Any suitable primary cooling medium602may be used including, for example, water or mixtures of water and glycol.

The heat absorbed by the primary cooling medium602is subsequently rejected at either a second coil614in an economizer mode or a heat exchanger616in an active mode. The primary cooling medium602is circulated through the primary coolant loop610and to either the second coil614or the heat exchanger616by a pump618. A diverter valve620selectively directs the pumped primary cooling medium602to either the second coil614or the heat exchanger616, depending upon the mode.

FIG.20shows the cooling system110in the economizer mode. The economizer mode is used, as with the passive modes discussed in the embodiments above, when the ambient air temperature is less than the temperature of the primary cooling medium602after it has absorbed the heat from the return air114(e.g., measured at a point in the primary coolant loop610after the cooling coil612). As with the embodiments discussed above, a predetermined temperature differential may be imposed to determine when the economizer mode or active mode is used. In the economizer mode, the diverter valve620directs the primary cooling medium602from the cooling coil612to the second coil614to cool the primary cooling medium602. Scavenger air118is directed across an outer surface of the second coil614by the scavenger fans120. The heat in the primary cooling medium602is then rejected from the primary cooling medium602and absorbed by the scavenger air118. The primary cooling medium602then returns to the cooling coil612. The cooling coil612and second coil614may be any suitable coil including finned tube coils or microchannel coils, for example. An expansion tank622may be located upstream of the pump618and after the cooling coil612.

FIG.21shows the cooling system110in the active mode. As with the active modes discussed above, the active mode of this embodiment is used when the ambient temperature is greater than the temperature of the primary cooling medium602after it has absorbed the heat from the return air114or within the predetermined temperature differential. In the active mode, the diverter valve620directs the primary cooling medium602from the cooling coil612to the heat exchanger616to cool the primary cooling medium602. The heat in the primary cooling medium602is then rejected from the primary cooling medium602and absorbed by a secondary cooling medium208of a secondary cooling system230. As discussed in the embodiments above, the secondary cooling system230may be any suitable cooling system including a direct expansion cooling system. The primary cooling medium602then returns to the cooling coil612.

Although this invention has been described in certain specific exemplary embodiments, many additional modifications and variations will be apparent to those skilled in the art in light of this disclosure. It is, therefore, to be understood that this invention may be practiced otherwise than as specifically described. Thus, the exemplary embodiments of the invention should be considered in all respects to be illustrative and not restrictive, and the scope of the invention to be determined by any claims supportable by this application and the equivalents thereof, rather than by the foregoing description.