Patent Application
20150047385
Air-source heat pumps are a common heating source in the southern United States and in many places around the globe. In many circumstances, heat pump systems can dramatically lower operating expenses when used in place of fossil fuels and resistive electric heating systems. However, air source heat pumps are often less favored in cold climates where fossil fuels or resistive electric heat are sometimes a more prevalent heating source. According to the U.S. Department of Energy's Energy Information Administration (EIA) this has resulted in higher costs for many consumers over the last decade as fossil fuel prices have increased an average of about 12 percent annually while electricity costs have increased at an average annual rate of only 3 percent for the same.
Heat pumps operate at a disadvantage in cold weather because as ambient air temperatures drop, most air-source heat pumps are often unable to satisfy the heat load requirement of the enclosed space to be heated. This is because air source heat pumps collect heat from the outside air and move it inside the enclosed space. However, as the outside air temperature decreases, heat is lost from the enclosed space at an ever-increasing rate while the heat pump must work ever harder to collect enough heat fast enough to maintain the indoor temperature. Thus, if the outside air becomes too cold, the air temperature in the enclosed space declines steadily as the air source heat pump struggles to collect enough heat causing the heated supply air entering the enclosed space to become steadily cooler. The supply air temperature can then begin to feel uncomfortable as it sinks below the body temperature of the occupants. If the heat demand cannot be maintained by the heat pump, electric resistance heat, i.e. strip heaters, fossil fuel furnaces or some other supplemental heat source must be engaged to introduce supplemental heat into the supply air to maintain the heat demand. However, using supplemental heat often considerably increases heating costs.
To collect and move heat into (or out of) an enclosed space, heat pump systems use a refrigerant or working fluid to move thermal energy along a circulation loop. A compressor raises the temperature and pressure of the refrigerant delivering it to a condensing unit. Heat is dissipated from the condensing unit causing the refrigerant to condense and change phase from a hot high-pressure gas to a warm high-pressure liquid. The pressurized and partially cooled refrigerant is then delivered to a metering device where its pressure is reduced as it enters an evaporating unit. Upon passing through the metering device and entering the evaporating unit, part of the working fluid changes phase from a warm high-pressure liquid to a two phase mixture of cool low pressure gas and liquid. During this phase change, some of the warm liquid condensate from the condensing unit quickly boils away (or “flashes”) to a gas thereby absorbing enough heat to cool the remaining liquid while the remainder of the working fluid remains a liquid in the evaporating unit. The remaining liquid evaporates by absorbing heat from an external medium outside the evaporator such as air, the ground, a supply of fluid such as water, or some other heat source. The evaporated refrigerant reenters the compressor, and the cycle is repeated continuously during normal operations.
In most residential settings, an air source heat pump system can either heat or cool an enclosed space by selectively controlling the flow of refrigerant through a series of valves and heat exchangers. Such systems are reversible in that depending on the situation, an indoor and outdoor heat exchanger can alternately operate either as an evaporator or a condenser depending on whether the system is operating in a heating mode or cooling mode. For example, in hot weather, an outdoor heat exchanger operates as a condenser dissipating excess heat of condensation into the ambient air while an indoor heat exchanger cools by absorbing heat of evaporation from the air inside the enclosed space. During the cooler months, the system is reversed causing the excess heat of condensation to be dissipated from the indoor heat exchanger to heat the enclosed space while the outdoor heat exchanger collects heat from the outside air to evaporate the refrigerant liquid entering the outdoor heat exchanger. This reversibility is provided in part because the substantial pressure reduction only occurs when the heat exchanger immediately downstream from a metering device is operating as an evaporator. When flowing in the opposite direction, the working fluid bypasses the pressure reduction elements within the metering device thus passing through the device without any substantial reduction in pressure.
However, the efficiency of an air-source heat pump typically drops off quickly as outdoor air temperatures drop below 35 degrees Fahrenheit. Heat losses from the enclosed space due to convection, conduction, radiation, and the like require that energy must be pumped back into the enclosed space at a particular rate to maintain a preset indoor temperature (usually set by a thermostat or similar control device). Therefore, working fluid in the outside heat exchanger must absorb at least this minimum amount of heat per hour (measured in British Thermal Units (BTU) per hour) by evaporating enough working fluid to carry the heat back to the compressor to repeat the cycle. This becomes increasingly more difficult as ambient air temperatures drop because collecting the necessary heat of evaporation rapidly enough from the low-temperature air to satisfy the heat demand becomes more and more difficult. The circulation of working fluid slows to accommodate the decreased rate of evaporation causing a reduction in heat transfer into the enclosed space which causes the indoor temperature to continue to fall until supplemental heating is engaged.
This failure to collect enough heat in the heating mode is worsened by the flash gas created as the working fluid enters the evaporator. As described above, some amount of working fluid immediately boils away upon entering the evaporating chamber because the working fluid cannot remain a liquid at a temperature higher than the boiling temperature corresponding to the pressure in the outdoor evaporator. The warm condensed liquid can no longer remain a liquid at the reduced pressures causing some part of the condensed liquid to evaporate to cool the remaining fluid in the liquid phase. However, this means that the heat used to evaporate the flash gas was not collected from the outside air but from the working fluid itself. This volume of gas therefore contributes nothing toward increasing the heat output of the system.
As the difference in temperature between the warm condensed liquid entering the metering device and the temperature inside the evaporator increases, so increases the quantity of working fluid that is immediately boiled away. The working fluid boiled off provides no benefit to the heat transfer and efficiency of the heat pump system because no heat energy is added to the working fluid to cause the phase change. Nonetheless, this volume of vaporized working fluid must still be recompressed by the compressor and passed to the indoor condenser. Thus the efficiency of the heat pump system is degraded because of the flash gas resulting from the phase change occurring in the metering device.
This unproductive working fluid can create a significant reduction in efficiency because heat pumps commonly use a fixed volume compressor to compress the working fluid at a predetermined number of cubic feet per minute (CFM). Therefore, the actual quantity of working fluid the system can circulate per minute is limited and generally does not vary while the heat carrying capacity, or “specific heat” of the working fluid being circulated may vary significantly. The specific heat of the working fluid is the amount of heat required to change its temperature. The heat required to cause the working fluid liquid to change phases from a liquid to a vapor is generally much higher than the heat required to raise the temperature of the working fluid already in the vapor phase. Therefore, it is generally advantageous to maintain the working fluid in a liquid phase as it enters the evaporating chamber to maximize the heat energy absorbed by the working fluid before entering the compressor. If all of the working fluid enters the evaporator as a fluid and is then evaporated to a gas, the compressor, and the heat pump system as a whole, operates at maximum capacity as all of the working fluid in the evaporating chamber contributes heat (BTUs) to the enclosed space.
Disclosed are various embodiments of a heat pump system configured to exchange heat between the cool fluid entering the evaporator from the metering device and the warm condensed working fluid entering the metering device from the upstream condenser. Thus the differential between these two separate fluid streams is reduced substantially to about zero degrees Fahrenheit and with it the non-productive gas that normally boils away when the working fluid passes through the metering device. Heat that would have otherwise been wasted by either generating unproductive vapor phase working fluid is conserved, or dissipated by other systems or devices incurring added cost and complexity, is reclaimed and used to further evaporate the liquid phase working fluid enhancing the efficiency of the heat pump system operating in the heating mode.
In one embodiment, a reversible heat pump system has a partitioned evaporator evaporating a working fluid in a first part and a second part using heat energy recovered from the working fluid in the first part. The system comprises a compressor coupled to a downstream condenser, and the partitioned evaporator has a first part exchanging heat between the working fluid from the upstream condenser at a first temperature and the working fluid from a metering device at a second temperature. The metering device receives the working fluid at a third temperature from the first part, the first part exchanging heat so that a second delta between the second and third temperatures is less than about 6 times larger than a first delta between the first and third temperatures. The second part is arranged downstream from the first part and upstream from the compressor such that the compressor, the condenser, the first part, the metering device, and the second part are coupled together to form a reversible closed refrigeration circuit for circulating the working fluid, the reversible closed refrigeration circuit operating in the heating mode to heat an enclosed space. In other similar embodiments, the second delta is less than about 5 times larger, less than about 3 times larger, or less than about 2 times larger than the first delta. In other embodiments, the first delta is greater than or equal to the second delta, at least about two times greater, at least about three times greater, or at least about five times greater than the second delta.
In another embodiment, a reversible heat pump system has a two-stage evaporator that transfers heat recovered from a working fluid in a first stage upstream from a metering device to the working fluid in a second stage downstream from the first stage. This embodiment of the system includes a condenser downstream from a compressor, and a multi-stage evaporator downstream from the condenser and upstream from the compressor. The multistage evaporator includes a first stage heat exchanger exchanging heat between a working fluid received from the condenser at an upstream temperature, and the working fluid from a metering device. The metering device receives the working fluid from the first part at a downstream temperature, and the first stage heat exchanger exchanges heat so that the downstream temperature of the working fluid is at least about 10 percent less than the upstream temperature during normal operations. The multistage evaporator further includes a second stage heat exchanger upstream from the compressor and downstream from the first stage, the second stage heat exchanger exchanging heat between a working fluid and an external medium. The reversible heat pump system is arranged such that the compressor, condenser, first heat exchanger, metering device, and second heat exchanger form a reversible closed refrigeration circuit for circulating the working fluid, the reversible closed refrigeration circuit operating in the heating mode to heat an enclosed space. In other similar embodiments, the downstream temperature of the working fluid is at least about 15 percent, at least about 20 percent, at least about 30 percent, at least about 50 percent, or at least about 90 percent less than the upstream temperature during normal operations.
In yet another embodiment, a reversible heat pump system is disclosed having a two-stage evaporator that provides a metering device between a first stage for evaporating a working fluid by reclaiming waste heat energy from the working fluid, and a second stage for evaporating the remaining working fluid. The two-stage evaporator has a first stage upstream from a metering device and downstream from a condenser, and a second stage downstream from the first stage. The first stage exchanges heat between the working fluid upstream from the first stage having an upstream temperature, and the working fluid upstream from the metering device having a downstream temperature. In this configuration, the first stage exchanges at least about 10 percent of heat exchanged by the second stage to reduce the difference between the upstream and downstream temperatures, and the second stage exchanges heat between the working fluid downstream from the first stage and an external medium. The reversible heat pump system also includes a compressor upstream from the condenser and downstream from the second stage of the two-stage evaporator, and is configured so that the compressor, condenser, first stage, metering device, and second stage form a reversible closed refrigeration circuit for circulating the working fluid, the reversible closed refrigeration circuit operating in the heating mode to heat an enclosed space. In related embodiments, the first stage exchanges at least about 15, at least about 20, at least about 30, or at least about 35 percent of heat exchanged by the second stage during normal operations. Similarly, in other related embodiments, the downstream temperature of the working fluid is at least about 10 percent less, or at least about 20 percent less than the upstream temperature during normal operations.
In any of the preceding embodiments, the second part or second stage may also be positioned downstream from the metering device and configured to receive working fluid directly or indirectly from both the metering device and from the first part or first stage. Similarly, for any of the preceding embodiments, the first part or first stage may exchange heat using any one of the following: a tube-in-tube heat exchanger, a coaxial coil heat exchanger, a plate heat exchanger, a shell-and-tube heat exchanger, or any other suitable heat exchanger. Also, for any of the preceding embodiments the second stage or second part may exchange heat with an external medium, such as any one of the following: ambient air, a liquid, or earth, or any suitable combination thereof.
Further forms, objects, features, aspects, benefits, advantages, and embodiments will become apparent from the detailed description and drawings provided herewith.
As noted above, included herein are various embodiments of a reversible heat pump system operating in the heating mode and configured to exchange heat between a relatively cool working fluid entering the outdoor evaporator from the metering device, and the relatively warm condensed working fluid entering the expansion device. In exchanging heat between these two separate flow paths, the difference in temperature between the fluid in the two flow paths is reduced, perhaps to substantially zero degrees Fahrenheit. This has the two-fold result of reducing or eliminating most if not all of the unproductive vapor created when the warm condensate enters the lower pressure evaporator downstream from the metering device while also recovering heat from the warm condensate that would otherwise be unused to aid in the evaporation of the liquid working fluid in the evaporator. Thus the refrigerant working fluid in the evaporator is arranged to absorb the maximum available heat optimizing overall system capacity.
The efficiencies that may be gained can be illustrated in nonlimiting example where a S-ton air-source heat pump utilizes a working fluid consisting of a mixture of difluoromethane (CH2F2, also known as R-32) and pentafluoroethane (CHF2CF3, also known as R-125), often mixed in equal parts and referred to by the American Society of Heating, Refrigerating, and Air Conditioning Engineers (ASHRAE) as R-410A. One example of a virtually available heat pump with similar specifications is the SSZ140361B 3-ton (36,000 BTU per hour) heat pump commercially available from the Goodman Manufacturing Company, L.P, of Houston, Tex.
For purposes of the example, the system is located in a region of the northern hemisphere, such as in Maine, Wisconsin, Minnesota, or North Dakota in the United States where the ambient (i.e. outdoor air) temperature may drop to 8 degrees Fahrenheit or less, sometimes much less, during the winter months. Similar climates appear in other regions such as northern Europe, northern parts of Asia, or in corresponding latitudes in the Southern Hemisphere as well.
While the outdoor temperature may be 8 degrees Fahrenheit or less, the indoor air temperature inside the enclosed space may be, for example, 75 degrees. Therefore the working fluid in a reversible heat pump system having a conventional evaporator operating in the heating mode can have a working fluid liquid downstream from the condenser and upstream from the metering device with a temperature of about 75 degrees. This temperature corresponds to about the room temperature inside the enclosed space. The difference, then, between the upstream and downstream temperature on both sides of the metering device is about 63 degrees. This temperature difference, as noted above, results in flash gas forming in the evaporator and a corresponding reduction in overall system efficiency. A greater difference between these two temperatures results in more flash gas causing a further reduction in overall system efficiency.
As noted above, heat pumps attempt to move heat into the enclosed space at a particular thermal transfer rate measured in BTUs per hour. On a cold day, the enclosed space loses heat at a particular rate that the heat pump must equal or exceed in order to maintain the desired temperature in the enclosed space. According to the Thermodynamic Properties of DuPont™ Suva® 410A Refrigerant (R-410A) available from E. I. du Pont de Nemours and Company, Wilmington, Del. and hereby incorporated herein by reference (pertinent parts of which are included herein as Appendix A), the enthalpy, or the total energy or total heat content, of R-410A vapor at 8 degrees Fahrenheit is 119.2 BTUs per pound. The condensed liquid upstream from the evaporator metering device is 75 degrees Fahrenheit and therefore has an enthalpy of 41.9 BTUs per pound, thus creating a difference of 77.3 BTUs per pound. As noted above, the compressor commonly compresses a fixed volume of working fluid vapor per minute making the density of the vapor an important aspect. Flash gas forms with a high specific volume (low density). Low density working fluid entering the compressor means the compressor must compress a greater volume of working fluid to collect enough working fluid in the condenser. Given that the compressor can only compress a fixed quantity of working fluid per minute, this means that the compressor must work longer to move the necessary heat into the enclosed space if the vapors in the evaporator having a relatively low density (relatively high specific volume).
Continuing the example from above, at 8 degrees Fahrenheit, the density of R-410A is 1.2255 pounds per cubic foot. Multiplying 1.2255 pounds per cubic foot by a common 3-ton compressor capacity of 2.8512 cubic feet per minute, the result is that 3.4941 pounds per minute of working fluid is circulated at 8 degrees Fahrenheit. Therefore the compressor handling 3.4941 pounds per minute of working fluid vapor carrying 77.3 BTUs per pound will transfer about 16,205 BTUs of heat into the enclosed space over the course of an hour. This result correlates closely with the capacity listed in the Product Specifications for the SSZ140361B heat pump published by the Goodman Manufacturing Company, L.P. and incorporated herein by reference (pertinent excerpts of which are included herein as Appendix B) which states the capacity at 5 degrees Fahrenheit to be 15.3 MBh (or 15,300 BTUs per hour) and the capacity at 10 degrees Fahrenheit to be 17.3 MBh (17,300 BTUs per hour).
However, as previously discussed, the enthalpy of a refrigerant does not change as it passes through the expansion valve but some quantity of the working fluid is vaporized because no substance can remain a liquid at a temperature higher than the boiling temperature corresponding to its pressure. This vaporization results in the fluid cooling and a loss of enthalpy equal to the difference between the enthalpy of the working fluid at 75 degrees Fahrenheit and the working fluid at the evaporating temperature of 8 degrees Fahrenheit. According to the previously mentioned Thermodynamic Properties of DuPont™ Suva® 410A Refrigerant, the enthalpy of liquid R-410A at 75 degrees Fahrenheit is 41.9 BTUs per pound, while the enthalpy of liquid R-410A at 8 degrees Fahrenheit is 16.6 BTUs per pound making the difference 25.3 BTUs per pound.
As noted above, the latent heat required to vaporize 1.0 pound of liquid R-410A at 8 degrees Fahrenheit is 119.2 BTUs per pound. The quantity of the working fluid flashing to vapor is determined by the ratio of the difference between the heat of the liquid entering the metering device and the heat of the liquid in the evaporator divided by the heat required to vaporize a pound of working fluid. In this case, the 25.3 BTUs per pound difference in enthalpy between the warm condensate and the cool working fluid is divided by 119.2 BTUs per pound latent heat of vaporization at 8 degrees Fahrenheit, or about 0.2122 which is equal to 21.22 percent. The remaining 78.78 percent of the liquid working fluid entering the evaporator remains in the liquid phase. As noted above, only vaporization of the remaining 78.78 percent of the liquid provides any additional benefit in heating the enclosed space. The 21.22 percent of the working fluid that flashes to vapor must be compressed again and passed to the condenser having provided no additional heating effect. At extreme cold outdoor temperatures, this percentage can exceed 50 percent further reducing system efficiency.
The disclosed embodiments increase heat pump efficiency by using a multi-stage or partitioned evaporator. In the first part or first stage, the working fluid exchanges heat between the warm condensate upstream from the metering device, and the cooler liquid downstream from the metering device before it enters the second stage or second part of the evaporator. In the second stage or second part, the working fluid absorbs heat from the air, the ground, water, or any other heat source similar to the process described above. This partitioned evaporator operates to both substantially reduce or eliminate the formation of unproductive flash gas as the working fluid passes through the metering device by reducing the difference between the upstream temperature and downstream temperatures of the working fluid entering and leaving the metering device. By equalizing the upstream and downstream temperatures, the condensed liquid enters the second stage of the evaporator at about evaporating temperature resulting in little if any flash gas being created. Increased efficiencies can be achieved by reducing the temperature of the working fluid entering the metering device by about 10 percent or more, more preferably by about 15 percent or more, by about 20 percent or more, or by about 25 percent or more, and most preferably by over 90 percent or more thus capturing the maximum available heat that can be captured by the first stage.
The fluid to fluid exchange in the first part also operates to transfer heat into the evaporator by using the warm condensate as a source of evaporative heat. Some or all of the cool liquid passing out of the expansion device is warmed by the heat exchange with the upstream warm condensate causing evaporation of the cool liquid. Thus heat (enthalpy) in the working fluid that could not be radiated into the enclosed space from the condenser can be used to warm or evaporate the liquid in the evaporator. The heat transfer into the working fluid moving into the second part of the evaporator can create working fluid vapor having a lower specific volume (higher density) and may also result in a corresponding increase in pressure in the second stage of evaporator. This increased pressure means the compressor need not work as hard to raise the pressure of the vaporized working fluid, and the higher density of the vapor means that more working fluid is compressed with each cubic foot of vapor.
The same 3-ton heat pump system discussed above under the same conditions but with the addition of the partitioned evaporator recovering the heat of condensation results in a significant efficiency gain. In this case, there is substantially no difference between the upstream and downstream temperatures. Therefore, the enthalpy of the upstream working fluid entering the metering device will be substantially equal to the enthalpy of the liquid in the second part of the evaporator. Put another way, the temperature of the condensate entering the expansion device will be substantially equal to the temperature of the working fluid entering the second part of the partitioned evaporator from the first part.
As noted previously, the enthalpy of liquid R-410A working fluid at 8 degrees Fahrenheit is 16.6 BTUs per pound. Therefore, using a partitioned evaporator the compressor will compress vapor during normal operations carrying 102.6 BTUs per pound of heat (119.2 BTUs per pound−16.6 BTUs per pound) rather than the 77.3 BTUs per pound of heat carried by the vapor in an conventional evaporator from the previous example above. Compressing the 3.4941 pounds per minute in a three ton unit like the one mentioned above means the system will move about 21,500 BTUs every hour into the enclosed space. This is about a 5,300 BTUs per hour increase in capacity over the same system without the partitioned evaporator or about 33 percent capacity increase. Furthermore, the 5,300 BTUs per hour gain in capacity resulting from removal of substantially all flash gas downstream from the metering device can be further increased by the disclosed embodiments as this heat can be recovered and used to warm or evaporate the liquid refrigerant rather than being rejected and lost as it would be in many subcooling systems. Instead, the otherwise “waste” heat of condensation is reintroduced into the second stage of the evaporator to evaporate remaining liquid phase working fluid thus improving evaporator capacity for a total capacity increase of about 10,600 BTUs per hour an increase of about 65 percent in heating capacity. Thus, in this example, where R-410A working fluid is used, the heat pump with a partitioned evaporator operates at 8 degrees Fahrenheit at nearly the same efficiency as an unpartitioned evaporator operating at 34 degrees Fahrenheit (according to the Product Specifications for the SSZ140361B heat pump).
The disclosed embodiments operating in a reversible heat pump system reduce the temperature of the working fluid entering the metering device. By reducing the temperature of the upstream condensed liquid, the side effect can be achieved that the metering device receives a steady flow of working fluid in the liquid phase rather than a mixture of liquid and gas. Many metering devices meter the flow of working fluid using a sensing device which may be mechanical or electronic, to detect the temperature of the working fluid leaving the evaporator. The metering device then responds by opening when vapors leaving the evaporator are too hot, and closing when the vapors are too cold. Metering devices can then be calibrated according to the working fluid in use and the application to ensure working fluid in the liquid phase does not enter the compressor which can damage it.
When a two phased mixture of liquid and gas working fluid enters the metering device, the hot gases generally pass quickly through the evaporator and into the compressor. The temperature sensor at the evaporator outlet senses the high temperature of the passing gas and causes the metering device to open, reacting quickly as if a large heat load were suddenly present and allowing a surge of condensed liquid into the evaporator. However, almost immediately, the bubble of hot vapor passes, and the cooler evaporated vapor is sensed by the sensor causing the expansion device to quickly close again. This opening and closing can become a regular occurrence in some cases as the expansion valve opens and closes in a recurring pattern responding to feedback from the temperature sensor. Such a condition is sometimes referred to “a hunting expansion valve” condition and results in erratic performance, abnormal wear on the metering device, and inefficiencies in overall performance of the system.
Recovering heat from the condensed liquid upstream from the metering device as disclosed and shown in the illustrated embodiments can also have the effect of reducing or eliminating the hunting expansion valve condition by reducing the temperature of the condensed liquid causing any hot vapor upstream from the first stage to recondense to liquid before entering the metering device. However, in order to achieve this effect, only a very small amount of heat need be recovered. For example, using a partitioned evaporator as discussed above and disclosed in greater detail below, the upstream condensate enters the first stage heat exchanger as a saturated liquid at about 75 degrees Fahrenheit at about 233 pounds per square inch absolute pressure (PSIA) having about 41.9 BTUs per pound of heat (See Appendix A). Cooling the condensate even a relatively small amount, for example to 73 degrees Fahrenheit, would likely eliminate any vapor in the lines caused by an increase in temperature above 75 degrees or a reduction in pressure below 233 PSIA. As shown in Appendix A, R-410A working fluid at 73 degrees Fahrenheit has a heat capacity of 41.1 BTUs per pound. Thus the heat to removed from the working fluid in order to maintain 73 degrees is 41.9 BTUs per pound−41.1 BTUs per pound or 0.8 BTUs per pound. At 3.4941 pounds of working fluid compressed per minute by the previously mentioned three ton compressor, the heat removed per hour is equal to 0.8 BTUs per pound multiplied by 3.4941 pounds per minute multiplied by 60 minutes per hour which equals about 168 BTUs per hour. Using this heat to evaporate liquid working fluid in a partitioned evaporator as described herein yields a total recovery of about 338 BTUs per hour which is about a 2 percent increase from the 16,200 BTUs per hour previously calculated for an unpartitioned evaporator. It can be seen, as well, that the number of BTUs exchanged by the first part in the partitioned evaporator to achieve this result is about 2 percent of the heat exchanged by the second part of the system. Thus, removing the vapor in the liquid condensate upstream from the metering device is a side effect of operating the disclosed partitioned evaporator. However, the disclosed embodiments can also recover well in excess of ten times the amount required to avoid a hunting expansion valve condition. In some circumstances, the disclosed system under normal operating conditions could recover perhaps as much as 50 times the heat required to avoid such a condition.
Various devices for avoiding a hunting valve condition, as well as for removing and discarding the excess heat extracted from the warm condensate to reduce the temperature differential across the metering device (sometimes referred to as “subcooling”) can be found in some non-reversible air conditioning systems. In such systems it is generally advantageous to cool the condensed liquid and reject the waste heat before it enters the evaporator. This is done to avoid the creation of flash-gas while also keeping heat out of the evaporator in the cooling mode so that maximum heat absorption can occur in the evaporator to cool the enclosed space. Adding any heat to an evaporator configured for cooling reduces its ability to absorb additional heat from the air or liquid load. Other uses for subcooling to eliminate heat from a refrigeration circuit include using some or all of the extracted heat to superheat the refrigerant vapor entering the compressor suction inlet, or ensuring that liquid refrigerant does not enter the compressor.
Techniques for achieving the advantages of subcooling include the incorporation of a dedicated heat exchanger on the downstream side of the condenser prior to the expansion device. This heat exchanger may be configured to exchange heat between the warm condensate and an external medium such as the air, ground, or perhaps a liquid bath containing water, brine, or other cool fluids. Lastly, some systems achieve a high degree of temperature reduction in the liquid condensate using a powered secondary cooling system in a heat exchange relationship with the warm condensate. Such systems are often used in cryogenic cooling systems or low-temperature refrigeration systems such as in supermarket refrigerators and freezers. However, powered subcooling equipment creates additional complexity and cost both to install and operate making it prohibitively expensive for most residential and commercial applications.
As will be illustrated in further detail below, the disclosed embodiments are arranged and configured to have substantially no impact (positive or negative) on the performance of a reversible heat pump system operating in the cooling mode. In the cooling mode, the partitioned evaporator located outside the enclosed space reverses roles operating instead as a condenser and the condenser in the enclosed space operates as an unpartitioned evaporator. The partitioned evaporator operating as a condenser receives hot compressed vapor phase working fluid from the upstream compressor in the second part or second stage and rejects heat from the working fluid vapor into the secondary medium (for example, ambient air). The condensed liquid phase working fluid then exits from the second stage and enters the first stage but without undergoing a pressure reduction as the metering valve positioned downstream from the second stage and upstream from the first stage is not configured to cause a significant pressure reduction between the two stages as it would be if operated in the heating mode. The warm condensed fluid then passes through the first stage and flows into a second metering device that is configured to reduce the pressure and allow the condensed liquid to expand and cool before entering the indoor evaporator (previously operating as a condenser). Heat of evaporation is then collected from the load (e.g. the air within the enclosed space) causing evaporated vapors to reenter the compressor.
Operating the disclosed partitioned evaporator inside the enclosed space to evaporate a liquid working fluid in the cooling mode causes a negative effect on the performance of the reversible heat pump system operating in the cooling mode. This is because it is fundamental to the operation of an air conditioner, or a reversible heat pump operating in the cooling mode, to remove as much heat from the load (e.g. the enclosed space) as possible by maintaining a large temperature differential between the liquid in the evaporator and the load. The disclosed embodiments, on the other hand, operate to collect heat from the warm condensed liquid entering the metering device and transfer it to the evaporating liquid, thus having the opposite effect of introducing heat into the evaporator that is not from the load. This additional heat is commonly the result of work performed by the compressor and is also commonly heat rejected from the condensed liquid by subcooling systems. Rejecting rather than collecting this heat is advantageous in the cooling mode because introducing additional heat into the evaporator from any source other than the load degrades the evaporator's ability to cool the load (as opposed to an evaporator operating in the heating mode were adding heat to the evaporating liquid is advantageous, regardless of the source). Therefore, although using the disclosed embodiments may be advantageous other purposes besides those disclosed, they do not include increasing the efficiency of a reversible heat pump system operating in the cooling mode.
Reference will now be made to the embodiments illustrated in the drawings, and specific language will be used to describe the same. It will nevertheless be understood that no limitation of the scope of the invention is thereby intended. Any alterations and further modifications in the described embodiments and any further applications of the principles described herein are contemplated as would normally occur to one skilled in the art to which the disclosure relates. Several embodiments are shown in great detail, although it will be apparent to those skilled in the relevant art that some, less relevant features may not be shown for the sake of clarity.
Reference numerals in the following description have been organized to aid the reader in quickly identifying the drawings where various components are first shown. In particular, the drawing in which an element first appears is typically indicated by the left-most digit(s) in the corresponding reference number. For example, an element identified by a “100” series reference numeral will first appear in
Outdoor heat transfer unit 102 operates as an evaporator in the heating mode for collecting heat from both an external medium and from working fluid 113. As heat is absorbed by the outdoor heat transfer unit 102, the working fluid changes phase from a liquid to a gas carrying with it the latent heat of evaporation 117 collected from the external medium. To evaporate working fluid 113, outdoor heat transfer unit 102 is partitioned into a first part or first stage 103 upstream from a reversible metering device 106, and a second part or second stage 105 downstream from metering device 106 and first part 103. First part 103 receives warm working fluid 113 from indoor heat transfer unit 111 operating as a condenser and exchanges heat from this upstream fluid with working fluid 113 received from the relatively low pressure, cooler downstream flow of working fluid 113 exiting reversible metering device 106. As a result working fluid 113 entering first part 103 from metering device 106 is warmed by heat from the relatively warm condensate working fluid 113. Conversely, the relatively warm working fluid 113 received from the indoor heat transfer unit 111 is cooled as well as it exits first part 103 and enters metering device 106. Evaporated working fluid 113 in the vapor phase enters downstream compressor 107 via reversing valve 109 carrying vapor from outdoor heat transfer unit 102 thus completing the reversible closed refrigeration circuit for circulating working fluid 113.
The closed refrigeration circuit is also made possible by a number of fluid conduits or lines 112 for carrying working fluid 113 between the various components of system 100. Lines 112 couple the compressor 107, reversing valve 109, indoor heat transfer unit 111, metering device 110, first part 103, metering device 106, second part 105, and reversing valve 109 as illustrated thus completing the reversible refrigeration circuit. Other components may also be included in the closed refrigeration circuit as well although they are not shown
Warm, substantially liquid, working fluid 113 upstream from metering device 106 is thus in a heat exchange relationship with cooler, also substantially liquid, fluid downstream from the metering device 106 and first part 103. This fluid-to-fluid exchange through and along one or more separate flow paths within first part 103 causes a reduction in the upstream temperature of fluid entering reversible metering device 106 transferring this upstream heat into the fluid exiting the metering device 106 to increase the downstream temperature of the fluid exiting metering device 106. Some quantity of working fluid 113 liquid phase may therefore be converted to vapor as it passes through first part 103 before entering second part 105. Thus working fluid 113 entering second part 105 from first part 103 can include a mixture of liquid and vapor phase working fluid already at about evaporating temperature and pressure. As a result, little if any further pressure reduction across metering device 106 is required as working fluid 113 enters second part 105. The remaining substantially liquid working fluid is retained in second part 105 to absorb heat 117 from an external medium.
During normal operations of reversible heat pump system 100, that is once the system reaches its normal operating temperatures and pressures, as opposed to the initial transitory fluctuations that occur on startup and shutdown, the temperature of working fluid 113 passing through first part 103 before entering metering device 106 is preferably at least about 10 percent cooler, more preferably at least about 15 percent or at least about 20 percent cooler, and most preferably at least about 90 percent cooler than the working fluid 113 entering first part 103 from indoor heat transfer unit 111.
In another aspect, a relationship between a first temperature of the working fluid received from the upstream indoor heat transfer unit 111 (operating as a condenser), a second temperature of working fluid 113 passing from metering device 106 into first part 103, and a third temperature of working fluid 113 passing from first part 103 into metering device 106 can be considered as well. Using these temperatures, a first delta is defined as the difference between the first temperature and the third temperatures, that is the temperature reduction as working fluid 113 passes through first part 103. Similarly, a second Delta can be defined as the difference between the second and third temperatures, that is the difference in temperature of the working fluid reentering first part 103 from metering device 106, and the working fluid entering metering device 106 from first part 103. Therefore, it may be preferable for second delta to be less than about 6 times larger than a first delta. It may be more preferable for the second delta to be less than about 5 times larger, less than about 3 times larger, or less than about 2 times larger than the first delta. It may be most preferable for the first delta to be greater than or equal to the second delta, at least about two times greater, at least about three times greater, or at least about five times greater than the second delta.
In yet another aspect of the disclosed embodiments, the first stage exchanges less heat between the separate working fluid flow paths during normal operations than the second stage or second part. In one embodiment, the first stage exchanges at least about 10 percent of heat exchanged by the second stage, while in another embodiment the first stage exchanges at least about 15 percent of heat exchanged by the second stage, or at least about 20 percent of heat exchanged by the second stage. Greater heat exchange in the first stage is envisioned as well such as the first stage exchanging at least about 25, 35, or 40 percent of heat exchanged by the second stage.
The closed refrigeration circuit shown at 100 is said to be “reversible” because it includes a reversing valve 109 as well as reversible metering devices 110 and 106. Reversing valve 109 is capable of changing the direction of flow of compressed working fluid 113 through lines 112 effectively reversing the roles of indoor heat transfer unit 111 and outdoor heat transfer unit 102 depending on whether the system is operating in the heating or cooling mode. Reversible metering devices 110 and 106 augment this reversibility by only allowing a pressure drop to occur across the individual metering devices as the fluid flows in one direction but not the other. For example, in the heating mode, metering device 110 is configured to avoid any substantial reduction in the flow or pressure of working fluid 113. On the other hand, reversible metering device 106 is configured to meter the flow of the warm working fluid entering either the first part 103. However, in the cooling mode, when reversing valve 109 positioned as shown in
Considering the cooling mode illustrated in
No specific implementation of any of the dimensions or components should be inferred from
Similarly, working fluid 113 may be any fluid suitable for transferring heat through a closed circuit vapor compression cycle such as a reversible heat pump system like the one illustrated in
Also, first part 103 of outdoor heat transfer unit 102 may include any device suitable for a heat exchange between upstream working fluid entering metering device 106 and downstream working fluid exiting metering device 106. For example, various types of heat exchangers are envisioned providing a plurality of separate fluid flow paths for exchanging heat between the upstream and downstream fluid flows on both sides of metering device 106. Examples of commercially available heat exchangers suitable for first part 103 include, but are not limited to, a tube-in-tube heat exchanger, a coaxial coil heat exchanger, a plate heat exchanger, or a shell-and-tube heat exchanger.
Also envisioned, although less preferred, is a first part 103 having upstream and downstream flow paths separated by a space or gap creating a heat transfer path across the gap which may include a secondary medium such as air or an external liquid. The heat exchange within first part 103 between the upstream warm condensate and the downstream cool fluid may be further enhanced by causing the working fluid 113 to flow in opposite directions along separate flow paths within first part 103. This retrograde or countercurrent flow, can enhance heat transfer between the plurality of separate flow paths possibly allowing reduced size or increased efficiency. Similarly, it should be understood that first part 103 represents embodiments of heat exchangers or similar devices having more than two flow paths including embodiments having three, four, ten, or any number of flow paths for working fluid 113.
Second part 105 of outdoor heat transfer unit 102 may be configured to exchange heat between the working fluid evaporating or condensing within second part 105 and an external medium such as ambient air, a liquid such as water or brine, or the earth such as in a direct or indirect exchange geothermal installation. Examples of devices that may be included in second part 105 include various types of tube and fin heat exchangers configured to exchange heat with ambient air or another external liquid, or other types of heat exchangers such as those used to exchange heat between a working fluid or some other liquid such as a solution containing water circulating in a ground loop in a geothermal installation.
Reversible metering devices 110 and 106 illustrated in
Enclosed space 114 may include various arrangements of openings such as doors and windows 115 which may be open or closed. Examples of enclosed space 114 include, but are not limited to, an office building, a commercial building, a bank, a multi-family dwelling such as an apartment building, a single family residential home, a factory, an enclosed or enclosable entertainment venue, a hospital, a store, a school, a single or multi-unit storage facility, a laboratory, a vehicle, an aircraft, a bus, a theatre, a partially and/or fully enclosed arena, a shopping mall, an education facility, a library, a ship, or other partially or fully enclosed structure.
Illustrated in
In
The working fluid entering second inlet 214 flows along a second flow path 215 that is separate from first flow path 209, passing downstream along second flow path 215 finally exiting first part 203 at a second outlet 216. As illustrated in
The working fluid 113 passes downstream from second outlet 216 entering second heat exchanger or second part 205 at a third inlet 217. In the embodiment illustrated in
After working fluid 113 evaporates to the vapor phase, working fluid 113 exits second part 205 entering a downstream compressor as illustrated in
Other embodiments of second part 105 from
In some embodiments, divider 313 may be configured to control and regulate the quantity of working fluid 113 entering second inlet 314 versus directly entering second part 305. For example, divider 313 may be configured to allow all of the working fluid to enter first part 303 before entering second part 305. In another embodiment, divider 313 might be positioned at second outlet 316 to divide the flow of fluid downstream from first stage 303 entering the second stage 305 thus dividing the working fluid after the exchange of heat in first stage 303 rather than before as illustrated. Other arrangements are envisioned as well such as dividing both the flow entering first stage 303 and dividing the flow exiting first stage 303 as well, thus having two dividers, further allowing for control over the quantity of warm condensate entering from the upstream condenser as well as the quantity of cooler substantially liquid fluid entering the downstream second stage 305. In another embodiment, divider 313 may be positioned as shown equally dividing the working fluid amongst the multiple paths.
Working fluid 113 arrives in second part 305 downstream from first part 303 entering one or more conduits 219 configured similar to the embodiment shown in
As discussed with respect to preceding figures and in greater detail with respect to first part 103 and 203, first part 303 may include any suitable apparatus useful for exchanging heat between the fluid upstream and downstream from metering device 212. Examples include, but are not limited to, tube-in-tube heat exchangers, coaxial coil heat exchangers, plate heat exchangers, or shell-and-tube heat exchangers, as well as others. Similarly, as discussed with respect to the second part 105 and 205, the heat exchange between working fluid 113 passing through multiple conduits 219 in second part 305 may also occur by exchanging heat with any suitable external medium such as ambient air, water or other liquid, or the ground in the case of geothermal loop, and others.
A perspective view of one embodiment of two-stage evaporator 300 is illustrated in
First part 503 has similar inlets and outlets creating a similar heat exchange relationship between one or more flow paths like those shown in
In the embodiments shown at 500 in
It may also be apparent in
A partitioned evaporator may also be used in reversible heat pump systems using an external medium such as the ground, or a liquid circulating through a loop buried in the ground or a body of water, such as a geothermal or similar system. In
Also included is a partitioned evaporator 602 having a first part or first stage 503 and a second part or second stage 605. First part 503 operates, and is arranged as shown and described previously with respect to
Working fluid 113 then enters second part 605 at a second part inlet 607 and exits second part 605 at a second part outlet 610 to reenter compressor 107 as a vapor. In the illustrated embodiment, second part 605 transfers heat of evaporation into working fluid 113 through a fluid to fluid transfer of a secondary fluid 611 that enters second part 605 at a secondary fluid inlet 614 and exits second part 605 at a secondary fluid outlet 616. The pumping device 618 causes secondary fluid 611 to circulate along the secondary closed loop 612 so that heat of vaporization from secondary fluid 611 transfers to working fluid 113 as working fluid 113 and secondary fluid 611 pass along separate flow paths within second part 605. Working fluid 113 is evaporated from a liquid phase to a vapor phase by heat exchanged with secondary fluid 611. Secondary fluid 611 is then cooled while working fluid 113 is warmed evaporating liquid working fluid to a vapor phase. Secondary fluid 611 is then reheated by heat 117 an external medium 620. Secondary fluid 611 can be circulated through any external medium 620 providing heat 117 such as the ground, a pool or lake of water, or any other suitable medium providing sufficient heat.
In other embodiments of heat pump system 600, secondary fluid 611 may not be present. In such cases, working fluid 113 may pass through a second part 605 some or all of which is directly contained within external medium 620 (such as the earth) creating a direct exchange between working fluid 113 and heat 117. One embodiment of this type of system is sometimes referred to as a “direct exchange” geothermal system where the working fluid 113 flows through lines buried in the earth. Such a system may also benefit from the use of a two-part or multistage partitioned evaporator.
It should be noted that recitation of ranges of values herein are merely intended to serve as a shorthand method of referring individually to each separate value falling within the range, unless otherwise indicated herein, and each separate value is incorporated into the specification as if it were individually recited herein. All methods described herein can be performed in any suitable order unless otherwise indicated herein or otherwise clearly contradicted by context. The use of any and all examples, or exemplary language (e.g., “such as”) provided herein, is intended merely to better illuminate the disclosure and does not pose a limitation on the scope of the invention unless otherwise claimed. No language in the specification should be construed as indicating any non-claimed element as essential to the practice of the invention.
The detailed descriptions and illustrations included herein are to be considered as illustrative and not restrictive in character, it being understood that only some embodiments have been shown and described and that all changes and modifications that come within the spirit of the invention are desired to be protected. In addition, all references cited herein are indicative of the level of skill in the art and are hereby incorporated by reference in their entirety.
