Patent Grant
7225627
In a closed-loop vapor compression cycle, the heat transfer fluid changes state from a vapor to a liquid in the condenser, giving off heat, and changes state from a liquid to a vapor in the evaporator, absorbing heat during vaporization. A typical vapor-compression system includes a compressor for pumping a heat transfer fluid, such as a freon, to a condenser, where heat is given off as the vapor condenses into a liquid. The liquid flows through a liquid line to a thermostatic expansion valve, where the heat transfer fluid undergoes a volumetric expansion. The heat transfer fluid exiting the thermostatic expansion valve is a low quality liquid vapor mixture. As used herein, the term “low quality liquid vapor mixture” refers to a low pressure heat transfer fluid in a liquid state with a small presence of flash gas that cools off the remaining heat transfer fluid, as the heat transfer fluid continues on in a sub-cooled state. The expanded heat transfer fluid then flows into an evaporator, where the liquid refrigerant is vaporized at a low pressure absorbing heat while it undergoes a change of state from a liquid to a vapor. The heat transfer fluid, now in the vapor state, flows through a suction line back to the compressor. Sometimes, the heat transfer fluid exits the evaporator not in a vapor state, but rather in a superheated vapor state.
In one aspect, the efficiency of the vapor-compression cycle depends upon the ability of the vapor compression system to maintain the heat transfer fluid as a high pressure liquid upon exiting the condenser. The cooled, high-pressure liquid must remain in the liquid state over the long refrigerant lines extending between the condenser and the thermostatic expansion valve. The proper operation of the thermostatic expansion valve depends upon a certain volume of liquid heat transfer fluid passing through the valve. As the high-pressure liquid passes through an orifice in the thermostatic expansion valve, the fluid undergoes a pressure drop as the fluid expands through the valve. At the lower pressure, the fluid cools an additional amount as a small amount of flash gas forms and cools of the bulk of the heat transfer fluid that is in liquid form. As used herein, the term “flash gas” is used to describe the pressure drop in an expansion device, such as a thermostatic expansion valve, when some of the liquid passing through the valve is changed quickly to a gas and cools the remaining heat transfer fluid that is in liquid form to the corresponding temperature.
This low quality liquid vapor mixture passes into the initial portion of cooling coils within the evaporator. As the fluid progresses through the coils, it initially absorbs a small amount of heat while it warms and approaches the point where it becomes a high quality liquid vapor mixture. As used herein, the term “high quality liquid vapor mixture” refers to a heat transfer fluid that resides in both a liquid state and a vapor state with matched enthalpy, indicating the pressure and temperature of the heat transfer fluid are in correlation with each other. A high quality liquid vapor mixture is able to absorb heat very efficiently since it is in a change of state condition. The heat transfer fluid then absorbs heat from the ambient surroundings and begins to boil. The boiling process within the evaporator coils produces a saturated vapor within the coils that continues to absorb heat from the ambient surroundings. Once the fluid is completely boiled-off, it exits through the final stages of the cooling coil as a cold vapor. Once the fluid is completely converted to a cold vapor, it absorbs very little heat. During the final stages of the cooling coil, the heat transfer fluid enters a superheated vapor state and becomes a superheated vapor. As defined herein, the heat transfer fluid becomes a “superheated vapor” when minimal heat is added to the heat transfer fluid while in the vapor state, thus raising the temperature of the heat transfer fluid above the point at which it entered the vapor state while still maintaining a similar pressure. The superheated vapor is then returned through a suction line to the compressor, where the vapor-compression cycle continues.
For high-efficiency operation, the heat transfer fluid should change state from a liquid to a vapor in a large portion of the cooling coils within the evaporator. As the heat transfer fluid changes state from a liquid to a vapor, it absorbs a great deal of energy as the molecules change from a liquid to a gas absorbing a latent heat of vaporization. In contrast, relatively little heat is absorbed while the fluid is in the liquid state or while the fluid is in the vapor state. Thus, optimum cooling efficiency depends on precise control of the heat transfer fluid by the thermostatic expansion valve to insure that the fluid undergoes a change of state in as large of cooling coil length as possible. When the heat transfer fluid enters the evaporator in a cooled liquid state and exits the evaporator in a vapor state or a superheated vapor state, the cooling efficiency of the evaporator is lowered since a substantial portion of the evaporator contains fluid that is in a state which absorbs very little heat. For optimal cooling efficiency, a substantial portion, or an entire portion, of the evaporator should contain fluid that is in both a liquid state and a vapor state. To insure optimal cooling efficiency, the heat transfer fluid entering and exiting from the evaporator should be a high quality liquid vapor mixture.
The thermostatic expansion valve plays an important role and regulating the flow of heat transfer fluid through the closed-loop system. Before any cooling effect can be produced in the evaporator, the heat transfer fluid has to be cooled from the high-temperature liquid exiting the condenser to a range suitable of an evaporating temperature by a drop in pressure. The flow of low pressure liquid to the evaporator is metered by the thermostatic expansion valve in an attempt to maintain maximum cooling efficiency in the evaporator. Typically, once operation has stabilized, a mechanical thermostatic expansion valve regulates the flow of heat transfer fluid by monitoring the temperature of the heat transfer fluid in the suction line near the outlet of the evaporator. The heat transfer fluid upon exiting the thermostatic expansion valve is in the form of a low pressure liquid having a small amount of flash gas. The presence of flash gas provides a cooling affect upon the balance of the heat transfer fluid in its liquid state, thus creating a low quality liquid vapor mixture. A temperature sensor is attached to the suction line to measure the amount of superheating experienced by the heat transfer fluid as it exits from the evaporator. Superheat is the amount of heat added to the vapor, after the heat transfer fluid has completely boiled-off and liquid no longer remains in the suction line. Since very little heat is absorbed by the superheated vapor, the thermostatic expansion valve meters the flow of heat transfer fluid to minimize the amount of superheated vapor formed in the evaporator. Accordingly, the thermostatic expansion valve determines the amount of low-pressure liquid flowing into the evaporator by monitoring the degree of superheating of the vapor exiting from the evaporator.
In addition to the need to regulate the flow of heat transfer fluid through the closed-loop system, the optimum operating efficiency of the vapor compression system depends upon periodic defrost of the evaporator. Periodic defrosting of the evaporator is needed to remove icing that develops on the evaporator coils during operation. As ice or frost develops over the evaporator, it impedes the passage of air over the evaporator coils reducing the heat transfer efficiency. In a commercial system, such as a refrigerated display cabinet, the build up of frost can reduce the rate of air flow to such an extent that an air curtain cannot form in the display cabinet. In commercial systems, such as food chillers, and the like, it is often necessary to defrost the evaporator every few hours. Various defrosting methods exist, such as off-cycle methods, where the refrigeration cycle is stopped and the evaporator is defrosted by air at ambient temperatures. Additionally, electrical defrost off-cycle methods are used, where electrical heating elements are provided around the evaporator and electrical current is passed through the heating coils to melt the frost.
In addition to off-cycle defrost systems, vapor compression systems have been developed that rely on the relatively high temperature of the heat transfer fluid exiting the compressor to defrost the evaporator. In these techniques, the high-temperature vapor is routed directly from the compressor to the evaporator. In one technique, the flow of high temperature vapor is dumped into the suction line and the vapor compression system is essentially operated in reverse. In other techniques, the high-temperature vapor is pumped into a dedicated line that leads directly from the compressor to the evaporator for the sole purpose of conveying high-temperature vapor to periodically defrost the evaporator. Additionally, other complex methods have been developed that rely on numerous devices within the vapor compression system, such as bypass valves, bypass lines, heat exchangers, and the like.
In an attempt to obtain better operating efficiency from conventional vapor-compression systems, the refrigeration industry is developing systems of growing complexity. Sophisticated computer-controlled thermostatic expansion valves have been developed in an attempt to obtain better control of the heat transfer fluid through the evaporator. Additionally, complex valves and piping systems have been developed to more rapidly defrost the evaporator in order to maintain high heat transfer rates. While these systems have achieved varying levels of success, the vapor compression system cost rises dramatically as the complexity of the vapor compression system increases. Accordingly, a need exists for an efficient vapor compression system that can be installed at low cost and operated at high efficiency.
According to a first aspect of the present invention, a vapor compression system is provided that maintains high operating efficiency by feeding a saturated vapor into the inlet of an evaporator. As used herein, the term “saturated vapor” refers to a heat transfer fluid that resides in both a liquid state and a vapor state with matched enthalpy, indicating the pressure and temperature of the heat transfer fluid are in correlation with each other. Saturated vapor is a high quality liquid vapor mixture. By feeding saturated vapor to the evaporator, heat transfer fluid in both a liquid and a vapor state enters the evaporator coils. Thus, the heat transfer fluid is delivered to the evaporator in a physical state in which maximum heat can be absorbed by the fluid. In addition to high efficiency operation of the evaporator, in one preferred embodiment of the invention, the vapor compression system provides a simple means of defrosting the evaporator. A multifunctional valve is employed that contains separate passageways feeding into a common chamber. In operation, the multifunctional valve can transfer either a saturated vapor, for cooling, or a high temperature vapor, for defrosting, to the evaporator.
In one form, the vapor compression system includes an evaporator for evaporating a heat transfer fluid, a compressor for compressing the heat transfer fluid to a relatively high temperature and pressure, and a condenser for condensing the heat transfer fluid. A saturated vapor line is coupled from an expansion valve to the evaporator. In one aspect of the invention, the diameter and the length of the saturated vapor line is sufficient to insure substantial conversion of the heat transfer fluid into a saturated vapor prior to delivery of the fluid to the evaporator. In one preferred embodiment of the invention, a heat source is applied to the heat transfer fluid in the saturated vapor line sufficient to vaporize a portion of the heat transfer fluid before the heat transfer fluid enters the evaporator. In one aspect of the invention, a heat source is applied to the heat transfer fluid after the heat transfer fluid passes through the expansion valve and before the heat transfer fluid enters the evaporator. The heat source converts the heat transfer fluid from a low quality liquid vapor mixture to a high quality liquid vapor mixture, or a saturated vapor. Typically, at least about 5% of the heat transfer fluid is vaporized before entering the evaporator.
In one embodiment of the invention, the expansion valve resides within a multifunctional valve that includes a first inlet for receiving the heat transfer fluid in the liquid state, and a second inlet for receiving the heat transfer fluid in the vapor state. The multifunctional valve further includes passageways coupling the first and second inlets to a common chamber. Gate valves positioned within the passageways enable the flow of heat transfer fluid to be independently interrupted in each passageway. The ability to independently control the flow of saturated vapor and high temperature vapor through the vapor compression system produces high operating efficiency by both increased heat transfer rates at the evaporator and by rapid defrosting of the evaporator. The increased operating efficiency enables the vapor compression system to be charged with relatively small amounts of heat transfer fluid, yet the vapor compression system can handle relatively large thermal loads.
In yet another embodiment, heat transfer fluid enters the common chamber of the multifunctional valve as a liquid vapor mixture and generally follows a flow direction. By controlling the flow rate of the heat transfer fluid and the shape of the common chamber, its is possible to separate a substantial amount of the liquid vapor mixture into liquid and vapor so that heat transfer fluid exists the common chamber through an outlet as liquid and vapor, wherein a substantial amount of the liquid is separate and apart from a substantial amount of the vapor.
In one embodiment, the vapor compression system includes a compressor, a condenser, an evaporator, an XDX valve, and an expansion valve. In accordance with this embodiment, the flow of heat transfer fluid from the condenser to the evaporator can be switched to go through either the XDX valve or the expansion valve. Preferably, the vapor compression system includes a sensor that measures the conditions of ambient surroundings, that is, the area or space in which the conditions such as temperature and humidity are controlled or altered by vapor compression system. Upon determining the conditions of the ambient surroundings, the sensor then decides whether to direct the flow of heat transfer fluid to either the XDX valve or the expansion valve.
Another aspect of the invention provides a method of operating a vapor compression system, comprising: compressing a heat transfer fluid in a compressor; condensing the heat transfer fluid in a condenser; expanding the heat transfer fluid in an expansion device to form an expanded heat transfer fluid and supplying the expanded heat transfer fluid to an evaporator feed line, at least one of the expansion device, a diameter of the evaporator feed line, and a length of the evaporator feed line converting a significant amount of a liquid form of the expanded liquid heat transfer fluid to a high quality liquid vapor mixture; supplying the high quality liquid vapor mixture to an evaporator coil having a heat transfer surface, converting a portion of a liquid form of the high quality liquid vapor mixture to a vapor form within the evaporator coil; and returning the heat transfer fluid to the compressor.
In one embodiment of this aspect, at a fixed cooling load, the heat transfer surface of the evaporator coil is smaller than that required to obtain an equivalent evaporator capacity when the significant amount of the liquid heat transfer fluid is not converted from a liquid form to a high quality liquid vapor mixture.
In another embodiment of this aspect, at a fixed cooling load, the conversion of the significant amount of the liquid refrigerant from a liquid form to a high quality liquid vapor mixture allows for at least an equivalent evaporator capacity to be achieved using an decreased heat transfer fluid load when compared to the heat transfer fluid load required when the significant amount of the liquid heat transfer fluid is not converted from a liquid form to a high quality liquid vapor mixture.
In another embodiment of this aspect, operating at a fixed cooling load, the conversion of the significant amount of the liquid heat transfer fluid from a liquid form to a high quality liquid vapor mixture allows for at least an equivalent evaporator capacity to that achieved when the significant amount of the liquid heat transfer fluid is not converted from a liquid form to a high quality liquid vapor mixture and wherein a distributor is present between the evaporator feed line and the evaporator coil.
An embodiment of a vapor-compression system 10 arranged in accordance with one embodiment of the invention is illustrated in
The vapor compression system of the present invention can utilize essentially any commercially available heat transfer fluid including refrigerants such as, for example, chlorofluorocarbons such as R-12 which is a dicholordifluoromethane, R-22 which is a monochlorodifluoromethane, R-500 which is an azeotropic refrigerant consisting of R-12 and R-152a, R-503 which is an azeotropic refrigerant consisting of R-23 and R-13, and R-502 which is an azeotropic refrigerant consisting of R-22 and R-115. The vapor compression system of the present invention can also utilize refrigerants such as, but not limited to refrigerants R-13, R-113, 141b, 123a, 123, R-114, and R-11. Additionally, the vapor compression system of the present invention can utilize refrigerants such as, for example, hydrochlorofluorocarbons such as 141b, 123a, 123, and 124, hydrofluorocarbons such as R-134a, 134, 152, 143a, 125, 32, 23, and azeotropic HFCs such as AZ-20 and AZ-50 (which is commonly known as R-507). Blended refrigerants such as MP-39, HP-80, FC-14, R-717, and HP-62 (commonly known as R-404a), may also be used as refrigerants in the vapor compression system of the present invention. Accordingly, it should be appreciated that the particular refrigerant or combination of refrigerants utilized in the present invention is not deemed to be critical to the operation of the present invention since this invention is expected to operate with a greater system efficiency with virtually all refrigerants than is achievable by any previously known vapor compression system utilizing the same refrigerant.
In operation, compressor 12 compresses the heat transfer fluid, to a relatively high pressure and temperature. The temperature and pressure to which the heat transfer fluid is compressed by compressor 12 will depend upon the particular size of vapor compression system 10 and the cooling load requirements of the vapor compression system. Compressor 12 pumps the heat transfer fluid into discharge line 20 and into condenser 14. As will be described in more detail below, during cooling operations, second inlet 26 is closed and the entire output of compressor 12 is pumped through condenser 14.
In condenser 14, a medium such as air, water, or a secondary refrigerant is blown past coils within condenser 14 causing the pressurized heat transfer fluid to change to the liquid state. The temperature of the heat transfer fluid drops about 10 to 40° F. (5.6 to 22.2° C.), depending on the particular heat transfer fluid, or glycol, or the like, as the latent heat within the fluid is expelled during the condensation process. Condenser 14 discharges the liquefied heat transfer fluid to liquid line 22. As shown in
Those skilled in the art will recognize that vapor compression system 10 can be used in a wide variety of applications for controlling the temperature of an enclosure, such as a refrigeration case in which perishable food items are stored. For example, where vapor compression system 10 is employed to control the temperature of a refrigeration case having a cooling load of about 12000 Btu/hr (84 g cat/s), compressor 12 discharges about 3 to 5 lbs/min (1.36 to 2.27 kg/min) of R-12 at a temperature of about 110° F. (43.3° C.) to about 120° F. (48.9° C.) and a pressure of about 150 lbs/in2 (1.03 E5 N/m2) to about 180 lbs/in.2 (1.25 E5 N/m2)
In accordance with one preferred embodiment of the invention, saturated vapor line 28 is sized in such a way that the low pressure fluid discharged into saturated vapor line 28 substantially converts to a saturated vapor as it travels through saturated vapor line 28. In one embodiment, saturated vapor line 28 is sized to handle about 2500 ft/min (76 m/min) to 3700 ft/min (1128 m/min) of a heat transfer fluid, such as R-12, and the like, and has a diameter of about 0.5 to 1.0 inches (1.27 to 2.54 cm), and a length of about 90 to 100 feet (27 to 30.5 m). As described in more detail below, multifunctional valve 18 includes a common chamber immediately before the outlet. The heat transfer fluid undergoes an additional volumetric expansion as it enters the common chamber. The additional volumetric expansion of the heat transfer fluid in the common chamber of multifunctional valve 18 is equivalent to an effective increase in the line size of saturated vapor line 28 by about 225%.
Those skilled in the art will further recognize that the positioning of a valve for volumetrically expanding of the heat transfer fluid in close proximity to the condenser, and the relatively great length of the fluid line between the point of volumetric expansion and the evaporator, differs considerably from systems of the prior art. In a typical prior art system, an expansion valve is positioned immediately adjacent to the inlet of the evaporator, and if a temperature sensing device is used, the device is mounted in close proximity to the outlet of the evaporator. As previously described, such system can suffer from poor efficiency because substantial amounts of the evaporator carry a liquid rather than a saturated vapor. Fluctuations in high side pressure, liquid temperature, heat load or other conditions can adversely effect the evaporator's efficiency.
In contrast to the prior art, the inventive vapor compression system described herein positions a saturated vapor line between the point of volumetric expansion and the inlet of the evaporator, such that portions of the heat transfer fluid are converted to a saturated vapor before the heat transfer fluid enters the evaporator. By charging evaporator 16 with a saturated vapor, the cooling efficiency is greatly increased. By increasing the cooling efficiency of an evaporator, such as evaporator 16, numerous benefits are realized by the vapor compression system. For example, less heat transfer fluid is needed to control the air temperature of refrigeration case 36 at a desired level. Additionally, less electricity is needed to power compressor 12 resulting in lower operating cost. Further, compressor 12 can be sized smaller than a prior art system operating to handle a similar cooling load. Moreover, in one preferred embodiment of the invention, the vapor compression system avoids placing numerous components in proximity to the evaporator. By restricting the placement of components within refrigeration case 36 to a minimal number, the thermal loading of refrigeration case 36 is minimized.
While in the above embodiments of the invention, multifunctional valve 18 is positioned in close proximity to condenser 14, thus creating a relatively short liquid line 22 and a relatively long saturated vapor line 28, it is possible to implement the advantages of the present invention even if multifunctional valve 18 is positioned immediately adjacent to the inlet of the evaporator 16, thus creating a relatively long liquid line 22 and a relatively short saturated vapor line 28. For example, in one preferred embodiment of the invention, multifunctional valve 18 is positioned immediately adjacent to the inlet of the evaporator 16, thus creating a relatively long liquid line 22 and a relatively short saturated vapor line 28, as illustrated in
Preferably heat source 25 used to vaporize a portion of the heat transfer fluid comprises heat transferred to the ambient surroundings from condenser 14, however, heat source 25 can comprise any external or internal source of heat known to one of ordinary skill in the art, such as, for example, heat transferred to the ambient surroundings from the discharge line 20, heat transferred to the ambient surroundings from a compressor, heat generated by a compressor, heat generated from an electrical heat source, heat generated using combustible materials, heat generated using solar energy, or any other source of heat. Heat source 25 can also comprise an active heat source, that is, any heat source that is intentionally applied to a part of vapor compression system 10, such as saturated vapor line 28. An active heat source includes but is not limited to a source of heat such as heat generated from an electrical heat source, heat generated using combustible materials, heat generated using solar energy, or any other source of heat which is intentionally and actively applied to any part of vapor compression system 10. A heat source that comprises heat which accidentally leaks into any part of vapor compression system 10 or heat which is unintentionally or unknowingly absorbed into any part of vapor compression system 10, either due to poor insulation or other reasons, is not an active heat source.
In one preferred embodiment of the invention, temperature sensor 32 monitors the heat transfer fluid exiting evaporator 16 in order to insure that a portion of the heat transfer fluid is in a liquid state 29 upon exiting evaporator 16, as illustrated in
While the above embodiments rely on heat source 25 or the dimensions and length of saturated vapor line 28 to insure that the heat transfer fluid enters the evaporator 16 as a saturated vapor, any means known to one of ordinary skill in the art which can convert the heat transfer fluid to a saturated vapor upon entering evaporator 16 can be used. Additionally, while the above embodiments use temperature sensor 32 to monitor the state of the heat transfer fluid exiting the evaporator, any metering device known to one of ordinary skill in the art which can determine the state of the heat transfer fluid upon exiting the evaporator can be used, such as a pressure sensor, or a sensor which measures the density of the fluid. Additionally, while in the above embodiments, the metering device monitors the state of the heat transfer fluid exiting evaporator 16, the metering device can also be placed at any point in or around evaporator 16 to monitor the state of the heat transfer fluid at any point in or around evaporator 16.
Shown in
When a vapor compression system, such as vapor compression system 10, is in operation, heat transfer fluid is pumped through fast-action capillary tube 500 from inlet 505 to outlet 510, and gating valve 520 is opened to allow heat transfer fluid to exit from fast-action capillary tube 500. When a vapor compression system has ceased operation, or has been cycled off, gating valve 520 is closed to allow heat transfer fluid to fill up fast-action capillary tube 500. By allowing fast-action capillary tube 500 to fill up with heat transfer fluid, fast-action capillary tube 500 is able to immediately supply a unit, such as an evaporator, with a rush of heat transfer fluid in a liquid state. By being able to supply a unit, such as an evaporator, with a rush of heat transfer fluid in a liquid state, fast-action capillary tube 500 allows a vapor compression system to cycle on, or begin operation, rapidly.
Control line 33 is connected to an input 62 located on upper valve housing 44. Signals relayed through control line 33 activate the diaphragm within upper valve housing 44. The diaphragm actuates a valve assembly 54 (shown in
Shown in
An exploded perspective view of multifunctional valve 18 is illustrated in
While the above embodiments use a multifunctional valve 18 for expanding the heat transfer fluid before entering evaporator 16, any thermostatic expansion valve or throttling valve, such as expansion valve 42 or even recovery valve 19, may be used to expand heat transfer fluid before entering evaporator 16.
In one preferred embodiment of the invention heat source 25 is applied to the heat transfer fluid after the heat transfer fluid passes through expansion valve 42 and before the heat transfer fluid enters the inlet of evaporator 16 to convert the heat transfer fluid from a low quality liquid vapor mixture to a high quality liquid vapor mixture, or a saturated vapor. In one preferred embodiment of the invention, heat source 25 is applied to a multifunctional valve 18. In another preferred embodiment of the invention heat source 25 is applied within recovery valve 19, as illustrated in
As the high temperature heat transfer fluid nears common chamber 140, heat from the high temperature heat transfer fluid is transferred from the second passageway 123 to the common chamber 140 in the form of heat source 125. By applying heat from heat source 125 to the heat transfer fluid in common chamber 140, the heat transfer fluid in common chamber 140 is converted from a low quality liquid vapor mixture to a high quality liquid vapor mixture, or saturated vapor, as the heat transfer fluid flows through common chamber 140. Additionally, the high temperature heat transfer fluid in the second passageway 123 is cooled as the high temperature heat transfer fluid passes near common chamber 140. Upon traversing second passageway 123, the cooled high temperature heat transfer fluid exits second outlet 130 and enters condensor 14. Heat transfer fluid in common chamber 140 exits recovery valve 19 at first outlet 159 into saturated vapor line 28 as a high quality liquid vapor mixture, or saturated vapor.
While in the above preferred embodiment, heat source 125 comprises heat transferred to the ambient surroundings from a compressor, heat source 125 may comprise any external or internal source of heat known to one of ordinary skill in the art, such as, for example, heat generated from an electrical heat source, heat generated using combustible materials, heat generated using solar energy, or any other source of heat. Heat source 125 can also comprise any heat source 25 and any active heat source, as previously defined.
In one preferred embodiment of the invention, recovery valve 19 comprises third passageway 148 and third inlet 126. Third inlet 126 is connected to discharge line 20, and receives high temperature heat transfer fluid exiting compressor 12. A first gating valve (not shown) capable of terminating the flow of heat transfer fluid through common chamber 140 is positioned near the first inlet 124 of common chamber 140. Third passageway 148 connects third inlet 126 to common chamber 140. A second gating valve (not shown) is positioned in third passageway 148 near common chamber 140. In a preferred embodiment of the invention, the second gating valve is a solenoid valve capable of terminating the flow of heat transfer fluid through third passageway 148 upon receiving an electrical signal.
In accordance with the invention, vapor compression system 10 can be operated in a defrost mode by closing the first gating valve located near first inlet 124 of common chamber 140 and opening the second gating valve positioned in third passageway 148 near common chamber 140. In defrost mode, high temperature heat transfer fluid from compressor 12 enters third inlet 126 and traverses third passageway 148 and enters common chamber 140. The high temperature heat transfer fluid is discharged through first outlet 159 of recovery valve 19 and traverses saturated vapor line 28 to evaporator 16. The high temperature heat transfer fluid has a temperature sufficient to raise the temperature of evaporator 16 by about 50 to 120° F. (27.8 to 66.7° C.). The temperature rise is sufficient to remove frost from evaporator 16 and restore the heat transfer rate to desired operational levels.
During the defrost cycle, any pockets of oil trapped in the vapor compression system will be warmed and carried in the same direction of flow as the heat transfer fluid. By forcing hot gas through the vapor compression system in a forward flow direction, the trapped oil will eventually be returned to the compressor. The hot gas will travel through the vapor compression system at a relatively high velocity, giving the gas less time to cool thereby improving the defrosting efficiency. The forward flow defrost method of the invention offers numerous advantages to a reverse flow defrost method. For example, reverse flow defrost systems employ a small diameter check valve near the inlet of the evaporator. The check valve restricts the flow of hot gas in the reverse direction reducing its velocity and hence its defrosting efficiency. Furthermore, the forward flow defrost method of the invention avoids pressure build up in the vapor compression system during the defrost system. Additionally, reverse flow methods tend to push oil trapped in the vapor compression system back into the expansion valve. This is not desirable because excess oil in the expansion valve can cause gumming that restricts the operation of the expansion valve. Also, with forward defrost, the liquid line pressure is not reduced in any additional refrigeration circuits being operated in addition to the defrost circuit.
It will be apparent to those skilled in the art that a vapor compression system arranged in accordance with the invention can be operated with less heat transfer fluid those comparable sized system of the prior art. By locating the multifunctional valve near the condenser, rather than near the evaporation, the saturated vapor line is filled with a relatively low-density vapor, rather than a relatively high-density liquid. Alternatively, by applying a heat source to the saturated vapor line, the saturated vapor line is also filled with a relatively low-density vapor, rather than a relatively high-density liquid. Additionally, prior art systems compensate for low temperature ambient operations (e.g. winter time) by flooding the evaporator in order to reinforce a proper head pressure at the expansion valve. In one preferred embodiment of the invention, vapor compression system heat pressure is more readily maintained in cold weather, since the multifunctional valve is positioned in close proximity to the condenser.
The forward flow defrost capability of the invention also offers numerous operating benefits as a result of improved defrosting efficiency. For example, by forcing trapped oil back into the compressor, liquid slugging is avoided, which has the effect of increasing the useful life of the equipment. Furthermore, reduced operating cost are realized because less time is required to defrost the vapor compression system. Since the flow of hot gas can be quickly terminated, the vapor compression system can be rapidly returned to normal cooling operation. When frost is removed from evaporator 16, temperature sensor 32 detects a temperature increase in the heat transfer fluid in suction line 30. When the temperature rises to a given set point, gating valve 50 and multifunctional valve 18 is closed. Once the flow of heat transfer fluid through first passageway 38 resumes, cold saturated vapor quickly returns to evaporator 16 to resume refrigeration operation.
Those skilled in the art will appreciate that numerous modifications can be made to enable the vapor compression system of the invention to address a variety of applications. For example, vapor compression systems operating in retail food outlets typically include a number of refrigeration cases that can be serviced by a common compressor system. Also, in applications requiring refrigeration operations with high thermal loads, multiple compressors can be used to increase the cooling capacity of the vapor compression system.
A vapor compression system 64 in accordance with another embodiment of the invention having multiple evaporators and multiple compressors is illustrated in
In operation, multiple compressors 80 feed heat transfer fluid into an output manifold 82 that is connected to a discharge line 84. Discharge line 84 feeds a condenser 86 and has a first branch line 88 feeding a first multifunctional valve 90 and a second branch line 92 feeding a second multifunctional valve 94. A bifurcated liquid line 96 feeds heat transfer fluid from condenser 86 to first and second multifunctional valves 90 and 94. Saturated vapor line 68 couples first multifunctional valve 90 with evaporator 72, and saturated vapor line 70 couples second multifunctional valve 94 with evaporator 74. A bifurcated suction line 98 couples evaporators 72 and 74 to a collector manifold 100 feeding multiple compressors 80. A temperature sensor 102 is located on a first segment 104 of bifurcated suction line 98 and relays signals to first multifunctional valve 90. A temperature sensor 106 is located on a second segment 108 of bifurcated suction line 98 and relays signals to second multifunctional valve 94. In one preferred embodiment of the invention, a heat source, such as heat source 25, can be applied to saturated vapor lines 68 and 70 to insure that the heat transfer fluid enters evaporators 72 and 74 as a saturated vapor.
Those skilled in the art will appreciate that numerous modifications and variations of vapor compression system 64 can be made to address different refrigeration applications. For example, more than two evaporators can be added to the vapor compression system in accordance with the general method illustrated in
A multifunctional valve 110 arranged in accordance with another embodiment of the invention is illustrated in
In yet another embodiment of the invention, the flow of liquefied heat transfer fluid from the liquid line through the multifunctional valve can be controlled by a check valve positioned in the first passageway to gate the flow of the liquefied heat transfer fluid into the saturated vapor line. The flow of heat transfer fluid through the vapor compression system is controlled by a pressure valve located in the suction line in proximity to the inlet of the compressor. Accordingly, the various functions of a multifunctional valve of the invention can be performed by separate components positioned at different locations within the vapor compression system. All such variations and modifications are contemplated by the present invention.
Those skilled in the art will recognize that the vapor compression system and method described herein can be implemented in a variety of configurations. For example, the compressor, condenser, multifunctional valve, and the evaporator can all be housed in a single unit and placed in a walk-in cooler. In this application, the condenser protrudes through the wall of the walk-in cooler and ambient air outside the cooler is used to condense the heat transfer fluid.
In another application, the vapor compression system and method of the invention can be configured for air-conditioning a home or business. In this application, a defrost cycle is unnecessary since icing of the evaporator is usually not a problem.
In yet another application, the vapor compression system and method of the invention can be used to chill water. In this application, the evaporator is immersed in water to be chilled. Alternatively, water can be pumped through tubes that are meshed with the evaporator coils.
In a further application, the vapor compression system and method of the invention can be cascaded together with another system for achieving extremely low refrigeration temperatures. For example, two systems using different heat transfer fluids can be coupled together such that the evaporator of a first system provides a low temperature ambient. A condenser of the second system is placed in the low temperature ambient and is used to condense the heat transfer fluid in the second system.
Another embodiment of a multifunctional valve 225 is shown in
A first inlet 244 (corresponding to first inlet 24 in the previously described embodiment) receives liquid feed heat transfer fluid from expansion valve 42, and a second inlet 245 (corresponding to second inlet 26 of the previously described embodiment) receives hot gas from the compressor 12 during a defrost cycle. In one preferred embodiment multifunctional valve 225 comprises first inlet 244, outlet 248, common chamber 246, and expansion valve 42, as illustrated in FIG. F. In one preferred embodiment, expansion valve 42 is connected with first inlet 244. The valve body 226 includes a common chamber 246 (corresponding to common chamber 40 in the previously described embodiment). Expansion valve 42 receives heat transfer fluid from the condenser 14 which then passes through inlet 244 into a semicircular well 247 which, when gating valve 229 is open, then passes into common chamber 246 and exits from the multifunctional valve 225 through outlet 248 (corresponding to outlet 41 in the previously described embodiment).
A best shown in
Insofar as operation of multifunctional valve 225 is concerned, reference is made to the previously described embodiment since the components thereof function in the same way during the refrigeration and defrost cycles. In one preferred embodiment, the heat transfer fluid exits the condenser 14 in the liquid state passes through expansion valve 42. As the heat transfer fluid passes through expansion valve 42, the heat transfer fluid changes from a liquid to a liquid vapor mixture, wherein the heat transfer fluid is in both a liquid state and a vapor state. The heat transfer fluid enters the first inlet 244 as a liquid vapor mixture and expands in common chamber 246.
In one preferred embodiment, the heat transfer fluid expands in a direction away from the general flow of the heat transfer fluid. As the heat transfer fluid expands in common chamber 246, the liquid separates from the vapor in the heat transfer fluid. The heat transfer fluid then exits common chamber 246. Preferably, the heat transfer fluid exits common chamber 246 as a liquid and a vapor, wherein a substantial amount of the liquid is separate and apart from a substantial amount of the vapor. The heat transfer fluid then passes through outlet 248 and travels through saturated vapor line 28 to evaporator 16. In one preferred embodiment, the heat transfer fluid then passes through outlet 248 and enters evaporator 16 at first evaporative line 328, as described in more detail below. Preferably, the heat transfer fluid travels from outlet 248 to the inlet of evaporator 16 as a liquid and a vapor, wherein a substantial amount of the liquid is separate and apart from a substantial amount of the vapor.
In one preferred embodiment, a pair of gating valves 229 can be used to control the flow of heat transfer fluid or hot vapor into common chamber 246. In refrigeration mode, a first gating valve 229 is opened to allow heat transfer fluid to flow through first inlet 244 and into common chamber 246, and then to outlet 248. In defrost mode, a second gating valve 229 is opened to allow hot vapor to flow through second inlet 245 and into common chamber 246, and then to outlet 248. While in the above embodiments, multifunctional valve 225 has been described as having multiple gating valves 229, multifunctional valve 225 can be designed with only one gating valve. Additionally, multifunctional valve 225 has been described as having a second inlet 245 for allowing hot vapor to flow through during defrost mode, multifunctional valve 225 can be designed with only first inlet 244.
In one preferred embodiment, multifunctional valve 225 comprises bleed line 251, as illustrated in
In one preferred embodiment, multifunctional valve 225 is dimensioned as specified below in Table A and as illustrated in
In one preferred embodiment, the heat transfer fluid enters common chamber 246 through first inlet 244 as a low quality liquid vapor mixture 270. Liquid vapor mixture 270 is in both a liquid state and a vapor state, wherein the liquid is suspended within the vapor. As used herein, the heat transfer fluid that is in a liquid state will be referred to as liquid 280 and the heat transfer fluid that is in a vapor state will be referred to as vapor 285. As the heat transfer fluid passes from the inlet 244 of common chamber 246 to the outlet 248 of common chamber 246, a portion of liquid 280 coalesces. As used herein, the term “coalesces” means to unite or to fuse together. Therefore, when the phrase “a portion of liquid 280 coalesces” is used, it is meant that a portion of liquid 280 becomes united with or fused together with another portion of liquid 280. As the heat transfer fluid enters common chamber 246, liquid 280 is arranged with liquid vapor mixture 270 as liquid droplets suspended in vapor 280. After the heat transfer fluid enters common chamber 246 as a liquid vapor mixture 270, the slower moving liquid 280 begins to coalesce and settle at bottom surface 252 of common chamber 246 while the faster moving vapor 285 is forced through outlet 248, as illustrated in
In one preferred embodiment, as heat transfer fluid travels through common chamber 246, a portion of liquid 280 within liquid vapor mixture 270 coalesces into larger droplets which exit through outlet 248 along with vapor 285. In one preferred embodiment, the larger droplets of liquid 280 coalesces into a stream of liquid 280, wherein the stream of liquid 280 exits through outlet 248 along with a stream of vapor 285, as illustrated in
Common chamber 246 is divided into a first portion 290 and a second portion 295. First portion 290 includes first inlet 244 and outlet 248. By including first inlet 244 and outlet 248, first portion is also the portion of common chamber 246 upon which heat transfer fluid must flow through upon entering common chamber 246, and therefore the portion of common chamber 246 wherein flow direction 265 generally resides. Flow direction 265 is the general direction the heat transfer fluid flows as the heat transfer fluid travels from first inlet 244 to second inlet 248, as illustrated by arrows in
In one preferred embodiment, the flow of heat transfer fluid is in a turbulent state upon entering first inlet 244, so that a portion of vapor 285 gets trapped in second portion 295, creating eddy 275 in common chamber 246, and more preferably in second portion 295 of common chamber 246. Eddy 275 is a current of heat transfer fluid that flows in a generally circular direction, as illustrated in
Common chamber 246 can comprise any one of a variety of geometrical configurations which allow a portion of liquid 280 to coalesce within common chamber 246 and separate from liquid 280. In one preferred embodiment, first inlet 244 is a distance N1 away from outlet 248 and a distance N2 from back wall 253, wherein the sum of N1 and N2 equals the length of common chamber 246, as illustrated in
In one preferred embodiment, inlet 244 is adjacent with back wall 253 and bottom surface 252 is located a distance N3 from outlet 248 and a distance N4 from inlet 244, as illustrated in
In one preferred embodiment, the flow rate upon which heat transfer fluid is forced through first inlet 244 is increased to facilitate the separation of liquid 280 from vapor 285 in liquid vapor mixture 270, which causes liquid 280 to coalesce. For example, in a vapor compression system having a compressor of size X, a condenser of size Y, an evaporator of size Z, and first inlet 244 having a diameter of D, if the flow rate is increased from A to B, liquid 280 will more readily separate from vapor 285 and coalesce. Preferably, the flow rate of heat transfer fluid is increased so that the heat transfer fluid entering common chamber 226 is in a turbulent flow. More preferably, the flow rate of heat transfer fluid is increased so that the heat transfer fluid entering common chamber 246 is at such a rate that Eddy 275 forms within common chamber 246, as illustrated in
Upon passing through inlet 244, common chamber 246, and outlet 248, the heat transfer fluid enters first evaporative line 328. Preferably, first evaporative line 328 is insulated. Heat transfer fluid then exits first evaporative line 328 and enters evaporative coil 21. Upon exiting evaporative coil 21, heat transfer fluid enters second evaporative line 330. Heat transfer fluid exists in second evaporative line 330 and evaporator 16 at temperature sensor 32.
Preferably, every element within evaporator 16, such as saturated vapor line 28, multifunctional valve 225, and evaporator coil 21, absorbs heat. In one preferred embodiment, as the heat transfer fluid passes through expansion valve 42, the heat transfer fluid is at a temperature within 20° F. of the temperature of the heat transfer fluid within the evaporator coil 21. In another preferred embodiment, the temperature of the heat transfer fluid in any element within evaporator 16, such as saturated vapor line 28, multifunctional valve 225, and evaporator coil 21, is within 20° F. of the temperature of the heat transfer fluid in any other element within evaporator 16. While the above embodiments were described in reference to multifunctional valve 225, any multifunctional valve described herein, can be used as well.
In one preferred embodiment, vapor compression system 410 includes a compressor 412, a condenser 414, an evaporator 416, an XDX valve 418, and a metering unit 449, as illustrated in
Compressor 412 is coupled to condenser 414 by a discharge line 420. XDX valve 418 includes first inlet 461, second inlet 462 and outlet 463. Metering unit 449 includes inlet 464 and outlet 465. First inlet 461 of XDX valve 418 and inlet 464 of metering unit 449 are coupled to condenser 414 by a bifurcated liquid line 422.
A saturated vapor line 428 couples outlet 463 of XDX valve 418 to inlet 455 of evaporator 416, and a suction line 430 couples the outlet of evaporator 416 to the inlet of compressor 412. A refrigerant line 456 couples outlet 465 of metering unit 449 to inlet 455 of evaporator 416. A temperature sensor 432 is mounted to suction line 430 and is operably connected to XDX valve 418 and metering unit 449. Temperature sensor 432 relays temperature information through a control line 433 to XDX valve 418 and through a second control line 434 to metering unit 449.
In accordance with one preferred embodiment, the flow of heat transfer fluid from condenser 414 to evaporator 416 can be directed to go through either XDX valve 418 or metering unit 449. Preferably, the flow of heat transfer fluid from condenser 414 to evaporator 416 can be directed to go through either XDX valve 418 or metering unit 449 based on the conditions of the ambient surroundings 470. Ambient surroundings 470 is the area or space in which the conditions, such as temperature and humidity, are controlled or altered by vapor compression system 410. For example, if vapor compression system 410 was an air conditioning unit, then ambient surroundings 470 would be defined by the area within a building or house being cooled by the air conditioning unit. Moreover, if vapor compression system 410 was a refrigeration unit, for example, then ambient surroundings 470 would be the area within a freezer or a refrigerator being cooled by the refrigeration unit.
In one preferred embodiment, a sensor 460 is located in ambient surroundings 470 and measures the conditions of ambient surroundings 470. Sensor 460 is any metering device known to one of ordinary skill in the art that can measure the conditions of ambient surroundings 470, such as a pressure sensor, a temperature sensor, or a sensor that measures the density of the fluid. Sensor 460 relays information through a control line 481 to metering unit 449 and through a second control line 483 to XDX valve 418. In this way, sensor 460 is able to direct the heat transfer fluid to run either through XDX valve 418 or metering unit 449 based upon the conditions of ambient surroundings 470.
In one preferred embodiment, sensor 460 is located in ambient surroundings 470 and measures the humidity of ambient surroundings 470. A desired humidity level is programmed into sensor 460. Upon determining the humidity of ambient surroundings 470, sensor 460 then decides whether to direct the flow of heat transfer fluid to either XDX valve 418 or metering unit 449 based upon the desired humidity level programmed into sensor 460. If the desired humidity level is less than the actual humidity of the ambient surroundings 470, sensor 460 directs the flow of heat transfer fluid to flow through metering unit 449 by closing first inlet 461, and by opening inlet 464. By directing the heat transfer fluid to flow through metering unit 449, vapor compression system 410 operates in what will be referred to as a conventional refrigeration cycle. When vapor compression system 410 operates in a conventional refrigeration cycle, the amount of humidity in the ambient surroundings 470 is decreased. If the desired humidity level is greater than the actual humidity of the ambient surroundings 470, sensor 460 directs the flow of heat transfer fluid to flow through XDX valve 418 by opening first inlet 461, and by closing inlet 464. By directing the heat transfer fluid to flow through XDX valve 418, vapor compression system 410 operates in what will be referred to as an XDX cycle. When vapor compression system 410 operates in an XDX cycle, the amount of humidity in the ambient surroundings 470 increases.
In one preferred embodiment, gating valves 471 and 474 are located at first inlet 461 and inlet 464, respectively, as illustrated in
In one preferred embodiment, sensor 460 decides whether to direct the flow of heat transfer fluid to either XDX valve 418 or metering unit 449 based upon the temperature of the ambient surroundings 470. A desired temperature level for the ambient surroundings 470 must first be programmed into sensor 460. Sensor 460 directs the flow of heat transfer fluid to flow through metering unit 449 by closing first inlet 461 and by opening inlet 464. By directing the heat transfer fluid to flow through metering unit 449, vapor compression system 410 operates in what will be referred to as a conventional refrigeration cycle. When vapor compression system 410 operates in a conventional refrigeration cycle, the load capacity of vapor compression system 410 is decreased. If the desired temperature level cannot be reached after a predetermined time interval, then sensor 460 directs the flow of heat transfer fluid to flow through XDX valve 418 by opening first inlet 461 and by closing inlet 464. By directing the heat transfer fluid to flow through XDX valve 418, vapor compression system 410 operates in what will be referred to as an XDX cycle. When vapor compression system 410 operates in an XDX cycle, the load capacity of vapor compression system 410 is increased.
Varying the load capacity of vapor compression system 410 allows vapor compression system 410 to be more accurately sized for cooling ambient surroundings 470. For example, if ambient surroundings 470 needs to be cooled in a range which varies from an average amount of ° C. to a maximum amount of ° C., vapor compression system 410 must be sized to cool ambient surroundings 470 by at least the maximum amount of ° C. so that vapor compression system 410 can achieve the desired temperature level even when the difference between the temperature level of the ambient surroundings 470 and the desired temperature level is the maximum amount of ° C. However, this means that vapor compression system 410 must be sized larger than required, since more often than not vapor compression system 410 need only cool ambient surroundings by the average amount of ° C. However, by varying the load capacity of vapor compression system 410, as described above, vapor compression system 410 can be sized so that it cools ambient surroundings by the average amount of ° C. when operating vapor compression system 410 in a conventional refrigeration cycle, and up to the maximum amount of ° C. when operating vapor compression system 410 in an XDX cycle.
While the above use of sensor 460 to direct the flow of heat transfer fluid to either XDX valve 418 or metering unit 449 has been described as being in response to the humidity level or the temperature level of the ambient surroundings, sensor 460 may direct the flow of heat transfer fluid to either XDX valve 418 or metering unit 449 in response to any variable or condition. Moreover, while the above use of vapor compression system 410 has required a sensor 460 to direct the flow of heat transfer fluid to either XDX valve 418 or metering unit 449, the flow may be manually directed to either XDX valve 418 or metering unit 449, or directed to either XDX valve 418 or metering unit 449 in any one of a number of ways known to one of ordinary skill in the art, for any one of a number of reasons.
In one preferred embodiment, discharge line 420 is coupled to both second inlet 462 of XDX valve 418 and condenser 414, to facilitate the defrosting of evaporator 416. Preferably, discharge line 420 is bifurcated so as to allow discharge line 420 to be simultaneously coupled to both second inlet 462 of XDX valve 418 and condenser 414, as illustrated in
In one preferred embodiment, vapor compression system 10 includes a turbulent line 600 before the inlet of evaporator 16, as illustrated in
Preferably, turbulent line 600 is position between the metering unit, such as multifunctional valve 18, 90, 94, 110 or 225, recovery valve 19, XDX valve 418, or any conventional metering unit used to meter the flow of heat transfer fluid upon entering evaporator. The placement, size, and spacing of ridges 610 to create a turbulent flow depends on the diameter and length of turbulent line 600 along with the flow rate of the heat transfer fluid and the type of heat transfer fluid being used, all which are factors that can be determined by one of ordinary skill in the art. In one preferred embodiment, the line connecting the metering unit to the inlet of evaporator 16, referred to herein as either the saturated vapor line or the refrigerant line, includes turbulent line 600. Preferably, a portion of saturated vapor line or refrigerant line includes turbulent line 600.
As known by one of ordinary skill in the art, every element of vapor compression system 10 described above, such as evaporator 16, liquid line 22, and suction line 30, can be scaled and sized to meet a variety of load requirements. In addition, the refrigerant charge of the heat transfer fluid in vapor compression system 10, may be equal to or greater than the refrigerant charge of a conventional system.
Another embodiment of the present invention provides a high operating efficiency vapor compression system including an evaporator having more than one circuit. When operated according to the method of the present invention, such a system dispenses with the need for a distributor to partition the heat transfer fluid to the multiple circuits of the evaporator without the accompanying large loss in evaporator capacity typically seen when a conventional system is operated without a distributor.
In many applications, it is preferred to distribute heat transfer fluid from the expansion device into the circuits of a multi-circuit evaporator coil. In such applications, it is important to distribute the heat transfer fluid equally to each circuit of the evaporator coil. If this is not done, one or more circuits of the evaporator can become starved of heat transfer fluid. In such a situation, the evaporator capacity is reduced.
In conventional systems having a multi-circuit evaporator, if a simple manifold divider is used to partition the heat transfer fluid flow into the multiple evaporator circuits, the circuits of the evaporator coil tend not receive equal amounts of heat transfer fluid. Such a situation is illustrated in
The up-feed manifold receives heat transfer fluid at an input situated below multiple outputs. The down-feed manifold receives heat transfer fluid at an input situated above multiple outputs. The side-feed manifold receives heat transfer fluid at an input situated above some of the outputs but below other outputs. In each configuration, heat transfer fluid flows along the path of least resistance from the manifold input to the manifold output. As illustrated in
Many conventional systems include a “distributor” in an attempt to evenly distribute heat transfer fluid from an expansion device to the coils of a multi-coil evaporator. Typically, a distributor includes a nozzle positioned to focus heat transfer fluid flow evenly into a dispersion cone. Output passages are spaced evenly around the cone to receive the heat transfer fluid.
As illustrated in
In the method of the invention, the expanded heat transfer fluid is converted to a high quality liquid vapor mixture before delivery to the evaporator. Example III shows the results of a test performed using such a method and also using the conventional method of operation, i.e. where the expanded heat transfer fluid is not converted to a high quality liquid vapor mixture before delivery to the evaporator. Despite the absence of a distributor, conversion of the expanded heat transfer fluid to a high quality liquid vapor mixture before delivery to the evaporator allowed the evaporator capacity to be maintained. This was the case even with a reduction in the heat transfer surface of the evaporator.
In another embodiment of the invention, the increased efficiency obtained when a vapor compression system is operated according to the method of the present invention allows for a reduction in the heat transfer fluid load used in the system.
In another embodiment of the invention, the “heat transfer surface” of the evaporator coil is smaller than the heat transfer surface of an evaporator coil, manufactured from the same material, required to obtain an equivalent evaporator capacity when a significant amount of the liquid heat transfer fluid is not converted from a liquid form to a high quality liquid vapor mixture. For example, for an evaporator coil manufactured from a material such as copper, having a given diameter and wall thickness, the length of the evaporator coil may be reduced if the vapor compression system is operated according to the method of the present invention. For the purposes of the present invention, the “heat transfer surface” is the area of the evaporator coil in contact with the heat transfer fluid.
Evaporator capacity and mass flow rate are the principal measures of performance of refrigerant evaporators. Evaporator capacity is defined as the work done in terms of heat transfer fluid vaporized per hour. The Mass Flow Rate is the mass of heat transfer fluid that moves through the evaporator coil to be vaporized. Evaporator capacity commonly takes into consideration the amount of heat transfer fluid flow, the amount of heat removed, and the heat transfer rate. The expansion device size, the amount of heat transfer fluid in the system and the compressor capacity are each often used to commercially identify the mass flow rate.
Evaporator capacity is viewed as:
Q=U*A*(ΔT(log mean)), where
The evaporator capacity, Q, through the heating surface of an evaporator is the product of three factors;
A(m2)—the heat transfer surface,
U(Wm−2K−1)—the overall heat transfer coefficient, and
ΔT(log mean)—the overall temperature driving force(log mean).
The temperature driving force is a function of the refrigerant properties, the amount of refrigerant, and the amount of heat absorbed. The Overall Heat Transfer Coefficient is a function of the design of the evaporator. Factors affecting the Overall Heat Transfer Coefficient (U) include:
Without further elaboration it is believed that one skilled in the art can, using the preceding description, utilize the invention to its fullest extent. The following examples are merely illustrative of the invention and are not meant to limit the scope in any way whatsoever.
A 5-ft (1.52 m) Tyler Chest Freezer was equipped with a multifunctional valve in a refrigeration circuit, and a standard expansion valve was plumbed into a bypass line so that the refrigeration circuit could be operated as a conventional vapor compression system and as an XDX refrigeration system arranged in accordance with the invention. The refrigeration circuit described above was equipped with a saturated vapor line having an outside tube diameter of about 0.375 inches (0.953 cm) and an effective tube length of about 10 ft (3.048 m). The refrigeration circuit was powered by a Copeland hermetic compressor having a capacity of about ⅓ ton (338 kg) of refrigeration. A sensing bulb was attached to the suction line about 18 inches from the compressor. The circuit was charged with about 28 oz. (792 g) of R-12 refrigerant available from The DuPont Company. The refrigeration circuit was also equipped with a bypass line extending from the compressor discharge line to the saturated vapor line for forward-flow defrosting (See
XDX System—Medium Temperature Operation
The nominal operating temperature of the evaporator was 20° F. (−6.7° C.) and the nominal operating temperature of the condenser was 120° F. (48.9° C.). The evaporator handled a cooling load of about 3000 Btu/hr (21 g cal/s). The multifunctional valve metered refrigerant into the saturated vapor line at a temperature of about 20° F. (−6.7° C.). The sensing bulb was set to maintain about 25° F. (13.9° C.) superheating of the vapor flowing in the suction line. The compressor discharged pressurized refrigerant into the discharge line at a condensing temperature of about 120° F. (48.9° C.), and a pressure of about 172 lbs/in2 (118,560 N/m2).
XDX System—Low Temperature Operation
The nominal operating temperature of the evaporator was −5° F. (−20.5° C.) and the nominal operating temperature of the condenser was 115° F. (46.1° C.). The evaporator handled a cooling load of about 3000 Btu/hr (21 g cal/s). The multifunctional valve metered about 2975 ft/min (907 km/min) of refrigerant into the saturated vapor line at a temperature of about −5° F. (−20.5° C.). The sensing bulb was set to maintain about 20° F. (11.1° C.) superheating of the vapor flowing in the suction line. The compressor discharged about 2299 ft/min (701 m/min) of pressurized refrigerant into the discharge line at a condensing temperature of about 115° F. (46.1° C.), and a pressure of about 161 lbs/in2 (110,977 N/m2). The XDX system was operated substantially the same in low temperature operation as in medium temperature operation with the exception that the fans in the Tyler Chest Freezer were delayed for 4 minutes following defrost to remove heat from the evaporator coil and to allow water drainage from the coil.
The XDX refrigeration system was operated for a period of about 24 hours at medium temperature operation and about 18 hours at low temperature operation. The temperature of the ambient air within the Tyler Chest Freezer was measured about every minute during the 23 hour testing period. The air temperature was measured continuously during the testing period, while the vapor compression system was operated in both refrigeration mode and in defrost mode. During defrost cycles, the refrigeration circuit was operated in defrost mode until the sensing bulb temperature reached about 50° F. (10° C.). The temperature measurement statistics appear in Table I below.
Conventional System—Medium Temperature Operation With Electric Defrost
The Tyler Chest Freezer described above was equipped with a bypass line extending between the compressor discharge line and the suction line for defrosting. The bypass line was equipped with a solenoid valve to gate the flow of high temperature refrigerant in the line. An electric heat element was energized instead of the solenoid during this test. A standard expansion valve was installed immediately adjacent to the evaporator inlet and the temperature sensing bulb was attached to the suction line immediately adjacent to the evaporator outlet. The sensing bulb was set to maintain about 6° F. (3.33° C.) superheating of the vapor flowing in the suction line. Prior to operation, the vapor compression system was charged with about 48 oz. (1.36 kg) of R-12 refrigerant.
The conventional vapor compression system was operated for a period of about 24 hours at medium temperature operation. The temperature of the ambient air within the Tyler Chest Freezer was measured about every minute during the 24 hour testing period. The air temperature was measured continuously during the testing period, while the vapor compression system was operated in both refrigeration mode and in reverse-flow defrost mode. During defrost cycles, the refrigeration circuit was operated in defrost mode until the sensing bulb temperature reached about 50° F. (10° C.). The temperature measurement statistics appear in Table I below.
Conventional System—Medium Temperature Operation With Air Defrost
The Tyler Chest Freezer described above was equipped with a receiver to provide proper liquid supply to the expansion valve and a liquid line dryer was installed to allow for additional refrigerant reserve. The expansion valve and the sensing bulb were positioned at the same locations as in the reverse-flow defrost system described above. The sensing bulb was set to maintain about 8° F. (4.4° C.) superheating of the vapor flowing in the suction line. Prior to operation, the vapor compression system was charged with about 34 oz. (0.966 kg) of R-12 refrigerant.
The conventional vapor compression system was operated for a period of about 24½ hours at medium temperature operation. The temperature of the ambient air within the Tyler Chest Freezer was measured about every minute during the 24½ hour testing period. The air temperature was measured continuously during the testing period, while the vapor compression system was operated in both refrigeration mode and in air defrost mode. In accordance with conventional practice, four defrost cycles were programmed with each lasting for about 36 to 40 minutes. The temperature measurement statistics appear in Table I below.
1)one defrost cycle during 23 hour test period
2)three defrost cycles during 24 hour test period
As illustrated above, the XDX refrigeration system arranged in accordance with the invention maintains a desired the temperature within the chest freezer with less temperature variation than the conventional systems. The standard deviation, the variance, and the range of the temperature measurements taken during the testing period are substantially less than the conventional systems. This result holds for operation of the XDX system at both medium and low temperatures.
During defrost cycles, the temperature rise in the chest freezer was monitored to determine the maximum temperature within the freezer. This temperature should be as close to the operating refrigeration temperature as possible to avoid spoilage of food products stored in the freezer. The maximum defrost temperature for the XDX system and for the conventional systems is shown in Table II below.
The Tyler Chest Freezer was configured as described above and further equipped with electric defrosting circuits. The low temperature operating test was carried out as described above and the time needed for the refrigeration unit to return to refrigeration operating temperature was measured. A separate test was then carried out using the electric defrosting circuit to defrost the evaporator. The time needed for the XDX system and an electric defrost system to complete defrost and to return to the 5° F. (−15° C.) operating set point appears in Table III below.
As shown above, the XDX system using forward-flow defrost through the multifunctional valve needs less time to completely defrost the evaporator, and substantially less time to return to refrigeration temperature.
A three door reach in freezer was set up in two configurations and tested to determine the ability of the freezer to meet defined acceptance criteria under each configuration. The tests were conducted using a Three-door Reach-In freezer powered by a Copeland compressor (part number KAKD-011E-CAV) and loaded with 24 ozs of R-404A refrigerant. The compression circuit used a FSE-1/2-ZP35 expansion valve. In the unmodified configuration, the system capacity was rated by the manufacturer at 4,280 BTU/hr and the evaporator capacity at 3,500 BTU/hr.
In the first (unmodified) configuration, the freezer was operated as a conventional vapor compression system, i.e. without the conversion of the heat transfer fluid to a high quality liquid vapor mixture before delivery to the evaporator. In this configuration, the evaporator coil consisted of a total of forty-two (42) passes of ⅜″ copper tubing. The evaporator coil was fed by a double feed through a distributor.
In the second (modified) configuration, the freezer was operated according to the method of the present invention, i.e. portions of the heat transfer fluid were converted to a high quality liquid vapor mixture before delivery to the evaporator. In this configuration, the evaporator coil consisted of a total of twenty-eight passes of ⅜″ copper tubing. The evaporator coil was fed directly by a double feed without a distributor.
The test conditions were those set by Underwriters Laboratories as per NSF-7, 6.2. The test requires that a freezer shall be capable of maintaining an air temperature of 0° F. (−18° C.) or less in all freezer compartment interiors under defined environmental conditions.
The testing criteria require that, prior to the start of the test, the freezer is allowed to establish thermal equilibrium according to the manufacturer's instructions or cycle on and off at least two full cycles at an ambient temperature of 73±3° F. (22±2° C.). The test must be conducted within a test chamber maintained under the following conditions for the duration of the test:
Air temperatures within the empty freezer compartment must be monitored using remote sensing devices (thermocouples) accurate to a ±1° F. (0.5° C.). The thermocouples must be positioned as close as possible to the following locations:
Prior to recording the air temperatures, the unit must be operated for two complete refrigeration cycles at the test chamber ambient conditions. The temperature at each thermocouple location must then be recorded at 5-minute intervals over a period of 4 hours.
The time during which the freezer's compressor(s) is operating must be monitored over the complete test duration, and the compressor percentage run time must be calculated for each compressor using the formula: Compressor percentage run time, R=d/D×100, where: “d” is the elapsed time that the compressor is operating during a whole number of cycles; and “D” is the total elapsed time during a whole number of cycles.
In order to meet the acceptance criteria, the temperature at each thermocouple location within each freezer compartment must not exceed 0° F. (−18° C.) during the 4-hour test period, and the compressor percentage run time must not exceed 80%.
As shown in Table IV, the conventional system achieved the acceptance criteria, having a compressor run time percentage of 75%. Table V shows that the XDX (modified) system, i.e. the system operated so that the heat transfer fluid was converted to a high quality liquid vapor mixture before delivery to the evaporator, also achieved the acceptance criteria, even though no distributor was included to equalize the delivery of heat transfer fluid to the evaporator and the heat transfer surface is smaller that in the freezer operated by the conventional (unmodified) method. In addition, the compressor percentage runtime for the XDX (modified) system was less than that of the conventional system.
Thus, it is apparent that there has been provided, in accordance with the invention, a vapor compression system that fully provides the advantages set forth above. Although the invention has been described and illustrated with reference to specific illustrative embodiments thereof, it is not intended that the invention be limited to those illustrative embodiments. Those skilled in the art will recognize that variations and modifications can be made without departing from the spirit of the invention. For example, non-halogenated refrigerants can be used, such as ammonia, and the like can also be used. It is therefore intended to include within the invention all such variations and modifications that fall within the scope of the appended claims and equivalents thereof.
This application is a continuation-in-part of U.S. patent application Ser. No. 10/129,339, filed May 2, 2002, now U.S. Pat. No. 6,951,117, which is a National Stage of PCT/US00/14648, filed May 26, 2000. PCT/US00/14648 is a continuation-in-part of P.C.T. application PCT/US00/00663, filed Jan. 11, 2000, which was published in English and designated the United States and a continuation-in-part of U.S. patent application Ser. No. 09/431,830, filed Nov. 2, 1999, now U.S. Pat. No. 6,185,958. The contents of these prior applications are incorporated by reference.
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| Number | Date | Country |
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| Number | Date | Country | |
|---|---|---|---|
| 20050257564 A1 | Nov 2005 | US |
| Number | Date | Country | |
|---|---|---|---|
| Parent | 10129339 | US | |
| Child | 10948446 | US | |
| Parent | PCT/US00/00663 | Jan 2000 | US |
| Child | 10129339 | US | |
| Parent | 09431830 | Nov 1999 | US |
| Child | PCT/US00/00663 | US |
