Double Hybrid Heat Pump Systems and Methods of Use and Operations

STATEMENT REGARDING FEDERALLY SPONSORED RESEARCH OR DEVELOPMENT

Not Applicable

CROSS-REFERENCE TO RELATED APPLICATIONS

[0001] This application claims priority to, and the benefit of, U.S. Provisional Patent Application No. 63382097 filed on November 2, 2022, and is a continuation-in-part of PCT Patent Application No. PCT/US2022/072074 filed on May 3, 2022, which claims priority to, and the benefit of, U.S. Provisional Patent Application No. 63183615 filed on May 3, 2021, and is related to U.S. Patent Application No. 17661818 filed on May 3, 2022, now U.S. Patent No. 11768018, the disclosure and teachings of each is incorporated herein by reference in its entireties.

BACKGROUND OF THE INVENTION

Field of the Invention

[0002] The present invention generally relates to heat pumps. More specifically, the invention relates to hybrid heat pumps that may be used in combination with hot water and forced air systems for heating and domestic water supply and cooling systems in new and retrofit applications.

Background Art

[0003] Residential, commercial, and industrial heat pump systems are a growing market worldwide because they can be high efficiency and may be used to replace less efficient heating and cooling systems. The growth in heat pumps has been primarily in new homes or existing residences that have forced hot air heating systems. Presently, converting homes that are heated by hot water is prohibitively expensive for the average homeowner. Retrofitting these homes are limited to several expensive options, which include for example:

1. Replacing every radiator or baseboard with larger radiators or panel radiators or baseboards that can operate at low heating supply temperatures (between 110- 120 degrees)

2. Replacing the hot water system with a forced hot air system, 3. Using two heat pumps to provide hot air and water at the same time, or

4. Adding several hydronic air handlers and ductwork to supplement the heat output from the existing heating system.

[0004] The cost of these options can range from $70,000-$100,000 for a single family residence, which is far too expensive for a typical homeowner. Additionally, the level of complexity in these retrofits can make these jobs undesirable to installers as well. The uptake of heat pump systems in the commercial and industrial markets has been slower than residential due, at least in part, to the additional complexity and labor requirements of these systems.

[0005] In addition, supplemental point of use heating and cooling is desired in many buildings due to inadequate zoning or multiple apartments inside the building and/or the use of geothermal district loop.

[0006] As such, there is a continuing need for heat pumps, systems, and methods that may be used in new installations and retrofit applications for both existed forced air and water heating systems with lower cost and higher performance for residential, commercial and industrial buildings.

BRIEF SUMMARY OF THE INVENTION

[0007] The present invention addresses the above noted needs by providing various double hybrid heat pump system embodiments that may be used in new installations as well as retrofit applications for building with forced air or hot water heating systems and geothermal loops. In various embodiments, the system may employ one or more split heads for point of use distribution of point of use heating and cooling to different locations in a building, which are referred to herein as a Split-Hybrid systems.

[0008] The double hybrid heat pumps with enhanced performance includes a compressor for compressing low-pressure vapor phase refrigerant to high-pressure vapor phase refrigerant, a refrigerant condensing heat exchanger to directly or indirectly produce and store heated water using heat from the high-pressure refrigerant. The condensed refrigerant then proceeds through a 4- way valve, or reversing element, until it reaches a liquid receiver where liquid refrigerant may be stored and used by the system to automatically adjust the amount of refrigerant moving through the system at a given time. The liquid refrigerant proceeds to a refrigerant cooling heat exchanger in which the high-pressure liquid refrigerant is further cooled by exchanging heat with a secondary fluid, such as air, that is at a lower temperature than was used to exchange heat in the condensing heat exchanger. The high pressure, cooled liquid refrigerant is passed through a subcooling refrigerant-to-refrigerant heat exchanger which transfers heat to the low pressure refrigerant vapor, both warming the vapor and further cooling the high pressure liquid refrigerant. The further cooled high-pressure liquid refrigerant is passed through an expansion device, e.g., expansion valve, to drop the pressure of the cooled liquid. The low-pressure cooled liquid or liquid/gas two phase mixture refrigerant exiting the expansion device is then provided to a refrigerant evaporating heat exchanger to vaporize the low-pressure liquid refrigerant. The low pressure vaporized refrigerant passes through refrigerant-to-refrigerant heat exchanger, superheating heating the vaporized refrigerant before it returns to the compressor inlet through the suction line.

[0009] In various embodiments, the water heated by the condensing heat exchanger may be provided to one or both of 1) a hot water tank for storage or for use as domestic hot water and 2) a hydronic heating loop that may serve a variety of uses. For example, in a hot water heating system, in the heating mode, the heat pump may make hot water for use as either domestic hot water or hydronic hot water for heating, as well as providing hot air from the refrigerant to air heat exchanger for heating. Whereas, in the cooling mode, the heat pump may provide hot water for domestic hot water and cool air for air conditioning. In the cooling mode, the condensing heat exchanger transfers heat from the high pressure vapor refrigerant to the water or other heat exchange fluid and high pressure refrigerant will transfer additional heat through the cooling heat exchanger that served as the evaporating heat exchanger in heating mode to add heat to the refrigerant. In the cooling mode, if the conditions permit, the system can prioritize hot water production and force the refrigerant to fully condense in a condensing heat exchanger and then use a favorable temperature difference to further subcool the refrigerant in one or more heat exchangers before it reaches the thermal expansion device, increasing the net efficiency of the system greatly.

[0010] The enhanced performance heat pump may be used to eliminate many external components, complicated controls, and the huge amount of labor required to retrofit a home from an existing hot water heating system to a ground or water source heat pump. The enhanced heat pump system may use a combination of hot water and a secondary, lower temperature fluid, e.g., air, to provide heat. For this secondary heat exchanger, enhanced performance heat pumps may use one or more heat exchangers, such as refrigerant to air coil heat exchangers, to provide heating or cooling in the building, which may employ mini-split type indoor heads or other types of split heat exchangers. An advantage of the mini-split indoor heads is that they can be easily distributed on multiple floors of a building and placed inside bedrooms for summer cooling with no new ductwork required.

[0011] Accordingly, the present disclosure addresses the continuing need for HVAC systems with improved cost and performance.

BRIEF DESCRIPTION OF THE DRAWINGS

[0012] The accompanying drawings are included for the purpose of exemplary illustration of various aspects of the present invention, and not for purposes of limiting the invention, wherein:

[0013] FIGS. 1A-1D depict exemplary schematic embodiments of the double hybrid heat pumps in various configurations and modes of operation.

[0014] FIGS. 2A-2E depict exemplary schematic embodiments of the double hybrid heat pumps in various configurations and modes of operation.

[0015] FIGS. 3A-3B show exemplary refrigeration heating and cooling cycles of various embodiments.

[0016] FIGS. 4A-4D show exemplary geothermal retrofits using the prior art systems and methods of the present invention.

[0017] In the drawings and detailed description, the same or similar reference numbers may identify the same or similar elements. It will be appreciated that the implementations, features, etc. described with respect to embodiments in specific figures may be implemented with respect to other embodiments in other figures, unless expressly stated, or otherwise not possible.

DETAILED DESCRIPTION OF THE INVENTION

[0018] Double hybrid heat pump systems 50, and methods of use, operation, and control of the present invention may be employed in various heating and water supply solutions in a structure 200.

[0019] FIGS. 1A-2E depict exemplary schematic embodiments of double hybrid heat pump systems 50. In those embodiments, a compressor 1 receives at an inlet and then compresses a low-pressure vapor phase refrigerant to high-pressure vapor phase refrigerant, which passes through an outlet and is provided via connection 2 to an inlet to a refrigerant condensing heat exchanger 3. The refrigerant passes through the condensing heat exchanger 3 and is cooled by a first cooling fluid, which depending upon the application may he water or another fluid that may be used elsewhere. In various embodiments, the first cooling fluid may be circulated through the condensing heat exchanger 3 from a tank 36 using a storage tank circulator pump 35 via connections from 33, 34, 37 and 38 to remove heat from the high- pressure vapor phase refrigerant. It will be appreciated that the high-pressure vapor phase refrigerant may be fully or partially condensed to a liquid at the outlet to the condensing heat exchanger 3. Where the refrigerant is partially condensed, a two phase refrigerant mixture will exit the refrigerant condensing heat exchanger 3 and be fully condensed and subcooled in a cooling heat exchanger as described herein.

[0020] The condensed refrigerant is provided from the outlet of the condensing heat exchanger 3 via connection 4 to a reversing element 15 via port 5 serving as an inlet. It will be appreciated by one skilled in the art that word “port” as used herein may describe access points to hardware and/or software. For example, a port may serve as an inlet or entry point to a device or an outlet or exit point from a device depending upon the direction of fluid flow, current flow, etc. Similarly, a connection may be a direct or indirect physical or logical connection between hardware and/or software.

[0021] In heating modes, such as depicted in FIGS. 1A&1B, the reversing element 15 may be configured to pass the condensed refrigerant via connection 7 to a second reversing element 61, which will be used to control the direction of refrigerant flow through a refrigerant to secondary fluid heat exchanger 11. In many applications, the refrigerant to secondary fluid heat exchanger 11 may be one or more refrigerant to air heat exchangers that employ one or more blowers to provide heated or cooled air to one or more spaces in a building.

[0022] In various embodiments, one or more liquid receivers, e.g., tanks, 80 and 19 are provided to accommodate for the differing amounts of liquid refrigerant that may be required in the system 50 depending upon the operational modes. The total amount of refrigerant in the system 50, referred to as the “refrigerant charge”, will be constant, but depending on the conditions and operating modes, the relative amounts of refrigerant circulating in the system 50 and stored in the liquid receiver 80 will vary. In some operating modes, most of the volume from the condensing heat exchanger 3 to the expansion device 17 or 52 will be filled with liquid refrigerant, while in others, a portion of the piping and heat exchanger volume will be occupied by either superheated gas or partially condensed saturated refrigerant. The liquid receivers 80 provide a buffer allowing for storage and relative inflows and outflows of liquid refrigerant as the operating conditions of the system 50 vary. Since the volume inside the liquid receiver is best used when it is generally filled with liquid, and a temperature change must occur for the refrigerant to be converted to liquid, whichever active heat exchanger is directly upstream of the liquid receiver 80 will be the primary condenser of the system 50.

[0023] As shown in FIGS. 1A-1B, a reversing element 61 may be employed to move the liquid receiver tank 80 relative to the refrigerant-secondary fluid heat exchanger 11. As shown in FIG 1A, when the reversing element 61 is in a first configuration, liquid refrigerant enters the liquid receiver 80 before heat exchanger 11. When the reversing element 61 is in a second configuration, as shown in FIG IB, the refrigerant enters heat exchanger 1 1 before the liquid receiver 80. In this configuration, heat exchanger 3 may serve as a desuperheater or partial condenser and heat exchanger 11 may serve as the primary condenser. Any excess liquid refrigerant that is condensed by the different configurations of reversing element 61 may be stored in the liquid receiver 80.

[0024] Heat exchanger 11 , which serves as a refrigerant cooling heat exchanger for further cooling the fully or partially condensed liquid using a second cooling fluid, e.g., air, to remove the heat from the refrigerant. The second cooling fluid may be used for other applications, such as providing heated air for heating the structure 200, or exhausted.

[0025] It will be appreciated by the skilled artisan that by maintaining the temperature of the secondary cooling fluid, e.g., air, below the temperature of the first cooling fluid, e.g., water, the refrigerant may be subcooled more than achievable with only the first cooling fluid and greater efficiency may be derived from the system 50 relative to the prior art.

[0026] The heat exchanger 11 is usually deployed inside the structure 200. In various embodiments, the compressor 1, condensing heat exchanger-storage tank 36, and heat exchanger 11 may be housed in the same physical unit or multiple units that may be deployed in proximity for ease of installation and maintenance. It will be further appreciated that heat exchangers and other devices employed in the present invention may include one or more stages that may be operated as a single unit or separately by those skilled in the art.

[0027] In various embodiments using air as the second fluid, one or more blowers 12 may be provided proximate the heat exchanger 11 that may be controlled to control the amount of heat being transferred in the heat exchanger 11. The blowers 12 may be connected to ductwork inside the structure 200 to enable heated and cooled air to be distributed in the structure 200 in the heating and cooling modes, respectively. Various control algorithms may be used to control the amount of heat extracted by the blowers 12 to control for human comfort by balancing the flowrate of the air, refrigerant temperature and secondary effects caused by the further cooling of the refrigerant. By using the relatively cool temperature of the return air to subcool the condensed liquid entering the heat exchanger 11 via port 10, more heat may be extracted from the refrigerant and increase overall efficiency. The cooler refrigerant is then able to absorb more heat at heat exchanger 27 per pound of refrigerant that passes through the system.

[0028] In the heating mode, such as depicted in FIGS. 1A-1B, the high-pressure cooled refrigerant exits cooling refrigerant- to-refrigerant heat exchanger 51 via port 18 serving as an outlet and is provided to an expansion device 17, e.g., expansion valve. A reversable filter dryer 14 may be deployed between the refrigerant-air heat exchanger and expansion device 17, or elsewhere is the system 50 to remove debris from the refrigerant. The expansion device 17 imparts a pressure drop on the high-pressure cooled refrigerant and outputs the refrigerant in a low-pressure cooled liquid state or a two-phase mixture as may be desired. The expansion device 17 may be connected to a liquid receiver tank 19 to store refrigerant if the amount of refrigerant needed for heating and cooling mode differs for an embodiment of the system 50. The term expansion device as used herein includes expansion valves as well as other types of expansion devices designed to induce a pressure drop in the system 50.

[0029] In FIG. 1 A embodiments, the refrigerant leaves the liquid receiver tank 19 via port 20 and enters a refrigerant to source heat exchanger 27, which, in the heating mode serves as an evaporating heat exchanger for the refrigerant.

[0030] The heat exchanger 27 is sometimes deployed outside the structure 200. The heat exchanger 27 may be embodied in various heat exchanger designs employing various heat exchanger media including gas, solid, or liquid, as is known in the art. For example, the heat exchanger 27 may be a geo-thermal heat exchanger in which heat is exchanged with solid ground and/or water in a well, or a refrigerant- air heat exchanger.

[0031] The low-pressure vaporized refrigerant enters and exits the heat exchanger 27 via connections 26 and 25, respectively, and travels to reversing element 15. In heating mode, as depicted in FIG 1A-1B, and 2A-2C, the low-pressure vaporized refrigerant enters and exits the reversing element 15 via ports 8 and 9, respectively.

[0032] For those who are familiar with the art, commonly the expansion devices 17 or 52 may be controlled to create 10 degrees of super heat as measured by temperature sensor 93 or positioned downstream of the expansion devices 17 or 52. The low-pressure vaporized refrigerant leaves port 9 as a superheated gas and travels to port 48 of refrigerant-to-refrigerant heat exchanger 51, where it will be heated by the relatively warm liquid refrigerant that enters at port 45. The amount of additional superheat will vary depending on the conditions, but in many cases, the gaseous refrigerant may be heated 10-25 degrees Fahrenheit. [0033] The temperature of this superheated gas may be measured by a temperature sensor 93 and a control method may be implemented to allow the expansion device 17 to increase the flow of refrigerant into the evaporator 27, raising the suction pressure on the inlet side of the compressor 1 and reducing superheat at temperature sensor 92 below the 10 degree standard. The superheat will be heated further to 10 degrees of superheat or greater via the refrigerant- to-refrigerant heat exchanger 51 before the low pressure gaseous refrigerant enters the suction/inlet side of the compressor 1 . The liquid receiver 80 upstream of the expansion device 17 creates a stable source of liquid refrigerant, generally devoid of uncondensed gas, which allows expansion device 17 to be controlled more accurately and operate more stably.

In cooling modes, as depicted in FIGS. 1C-1D and 2D-2E, reversing element 15 is set to pass the high-pressure liquid refrigerant from condensing heat exchanger 3 to refrigerant-source heat exchanger 27 via connection 5 to port 8. In the cooling mode, the refrigerant to source heat exchanger 27 serves as a refrigerant subcooling heat exchanger further cooling the liquid refrigerant input via connection 25 and outputting subcooled high pressure liquid refrigerant via connection 26.

[0034] Similar to in heating mode, a control algorithm may be employed to adjust the amount of heat extraction from heat exchanger 27. If heat exchanger 27 has a lower temperature than heat exchanger 3, it may be capable of subcooling the refrigerant to a substantial degree, which will increase the cooling capacity of the evaporating heat exchanger 11 compared with using heat exchanger 3 alone. More generally speaking, when the refrigerant cooling heat exchanger is used to exchange heat with a heat exchange media that is at a lower temperature than the water exchanging heat with the refrigerant in the refrigerant condensing heat exchanger, the overall efficiency of the system 100 may be improved.

[0035] In cooling mode, with proper regulation of the speed of pump 30 to the source/sink heat exchanger, the system 50 will be able to condense high-pressure refrigerant in heat exchanger 3 and then further subcool the liquid refrigerant in heat exchanger 27. The amount of subcooling will depend on the relative temperatures of the interacting fluids on heat exchanger 3 and 27, and the size and design of the heat exchangers. The refrigerant, being thoroughly condensed and subcooled, will be capable of extracting more heat from heat exchanger 11 and improving efficiency in cooling mode. If desired, the system may maximize the heat release into the hot water tank 36 via heat exchanger 3 by turning off circulator pump 30, eliminating heat exchange through the source heat exchanger 27. [0036] In cooling scenarios, the liquid refrigerant may be passed through the subcooling refrigerant-to-refrigerant heat exchanger 51, which will further subcool the high pressure liquid refrigerant prior to entering the expansion device 52. The additional subcooling provided by refrigerant-to-refrigerant heat exchanger 51 will increase the amount of liquid refrigerant exiting the expansion device 52 and the superheat of the refrigerant evaporated in heat exchanger 11.

[0037] In the cooling embodiments FIG 1C-1 D and 2D-2E, the refrigerant passes through a check valve 54, bypassing expansion device 17, to the refrigerant-to-refrigerant heat exchanger 51 which acts as a subcooling heat exchanger for the high pressure liquid refrigerant and superheater for the low pressure vapor refrigerant entering at port 48. Excess liquid refrigerant, if any, may collect in the liquid receiver tank 19 at port 20 and re-enter the circulation loop through the check valve 54. The circulating high-pressure subcooled liquid refrigerant passes through the expansion device 52 which imparts a pressure drop to the refrigerant yielding low-pressure cooled liquid refrigerant.

[0038] The low-pressure cooled liquid refrigerant is provided via port 13 to the heat exchanger 11 , which serves as the evaporating heat exchanger. The refrigerant is partially or fully evaporated in the heat exchanger 11 and exits via port 10. In various embodiments, blowers 12 circulate air cooled by the heat exchanger 11 throughout the structure 200 via the ductwork. However, other heat exchange media may be employed in heat exchanger 11 depending upon various factors, such as desired efficiency and/or uses of the energy being transferred from the refrigerant.

[0039] In the cooling mode, the low-pressure vaporized refrigerant exiting the heat exchanger 11 via port 10 returns to the low pressure side of the subcooling heat exchanger 51 after passing through the reversing element 15 through ports 7 and 9. In the subcooling heat exchanger 51 , the low pressure evaporated refrigerant is superheated to a higher temperature by subcooling the high pressure liquid refrigerant. This additional superheating of the low pressure evaporated refrigerant being provided to the inlet of the compressor 1, i.e., the suction line, may be used in a number of ways:

1. The suction pressure may be held constant and the superheat of the refrigerant will be further raised before entering the compressor 1. The higher temperature of the low pressure vapor refrigerant entering the compressor 1 will lead to a higher discharge superheat of the discharge high pressure vapor refrigerant out of the compressor 1. The higher temperature vapor may be used to produce higher temperature liquids in tank 36 via heat exchanger 3 without raising the discharge pressure or increasing the compression ratio of the compressor 1 , both of which tend to decrease the efficiency of the system 50.

2. The suction pressure may be raised to reduce the superheat of the low temperature, low pressure vapor refrigerant as it leaves heat exchanger 11, while the industry standard of 10 degrees or greater of superheat of the low pressure vapor refrigerant entering the compressor 1 may still be obtained from superheating low pressure vapor in heat exchanger 51.

[0040] FIGS. 1A-2E embodiments may be employed to provide simultaneously both hot water for heating and domestic hot water use and hot air for heating in heating modes and hot water for domestic hot water use and cool air in cooling modes. Where applications may require, such as in a building with dehumidification that calls for simultaneous heating and cooling, chilled air and hot water or domestic hot water may be produced simultaneously as shown in FIG 1C and 2D. The hot water for use in both modes may be provided via the tank 36. Domestic hot water may be provided via a water to water heat exchanger 41, often called an indirect heat exchanger, in tank 36, via ports 42 and 43. Alternatively, a double wall heat exchanger may be used and domestic hot water may be produced directly in heat exchanger 3 and stored in tank 36.

[0041] It will be appreciated that hot water production may be effectively bypassed or reduced and the heat pump system 50 may be used solely for providing hot air by reducing the flow from circulator 35 to heat exchanger 3 and by using the reversing valve 61 to reverse the flow through the liquid refrigerant to secondary fluid heat exchanger 11. As shown in FIG IB and 2B, heat exchanger 11 may be used as a condensing and/or subcooling heat exchanger in these heating configurations. System 50 by turning off the blower for heat exchanger 11, system 50 may be configured to produce only hot water, as shown in FIG 2C. In this way, system 50 is a highly adaptable system, able to distribute heat or cooling as needed through a variety of different outlets with enhanced efficiency.

[0042] Similarly, forced hot or cool air distribution to the structure 200 may be reduced, stopped, or bypassed, if only hot water production was desired. In various embodiments, the blowers 12 may be slowed or not operated in the FIGS. 1A-2E configurations, which will reduce the heat transfer in heat exchanger 11 and the flow of air through the ductwork. Alternatively, the heat exchanger 11 may be bypassed using bypass valves (not shown).

[0043] FIG. 1A-2E embodiments include several elements that make the double hybrid heat pump more efficient and easier to install than prior art units. For example, heat introduced by the compressor 1 may be extracted via the condensing heat exchanger 3, which may have a pump that is installed proximate the heat exchanger 3 or the tank 36. The heat exchanger 3 may be sized to achieve a range of operational scenarios. For example, in various embodiments, the heat exchanger 3 may be sized to transfer the heat from the refrigerant until the refrigerant is -100% liquid at a pressure and temperature that is very close to the temperature of the incoming fluid, recognizing that heat transfer performance tends to vary over the life of a heat exchanger. For example, the heat exchanger 3 may be designed to have a small approach temperature, so that the temperature of the water leaving the heat exchanger at port 34 and the temperature of the refrigerant condensing into liquid entering the heat exchanger 3 at port 2 will be within a few degrees, e.g., 2-3, of each other under full load conditions.

[0044] In addition, the blower speed, and hence the amount of heat transferred in refrigerant-air heat exchanger 11 may be modulated based on a feedback loop to target a specific final refrigerant temperature and/or discharge air temperature. When the refrigerant enters heat exchanger 11 as a condensed liquid, much of the heat from the refrigerant has already been removed by the condensing heat exchanger 3, and therefore the blower 12 may be operated at lower speeds, so to not blow a large volume of air into the building which may be unpleasant to the occupants. The blower 12 may be set to cool the refrigerant to a predetermined temperature above the incoming air temperature. For example, if the incoming air temperature is 60 degrees and the refrigerant saturation temperature is 120 degrees and the predetermined setpoint is 15 degrees above the entering air temperature, the blower 12 would modulate to set the leaving refrigerant temperature at port 13 to be 75 degrees (60+15), yielding 45 degrees of subcooling from the discharge saturation temperature.

[0045] The effect of using a subcooling heat exchanger, (heat exchanger 11 when heating and heat exchanger 27 when cooling) is significant. The extra subcooling improves the overall efficiency of the refrigeration cycle by putting a larger load on the evaporator. Furthermore, since the subcooling is done by the subcooling heat exchanger, the pressure and temperature of the refrigerant entering the condensing heat exchanger may be lower, lowering the work done by the compressor.

[0046] A similar effect happens in the evaporator. When the high-pressure refrigerant is subcooled by a colder fluid in the refrigerant cooling heat exchanger the refrigerant passes through the expansion device with less residual heat. As a result, the expansion device will throttle less and allow the refrigerant pressure to rise, increasing efficiency and decreasing the compressor power input.

[0047] Performance data collected from experimental testing shows that 30 degrees of additional subcooling improves heat extraction by the evaporator heat exchanger 27 by between 25-30%. In addition, this increase in performance also reduced compressor power consumption by 3%.

[0048] Under the heavy lift conditions that are common to air source heat pumps or ground source heat pumps under peak conditions, it is possible to achieve subcooling in excess of 70 degrees. In FIG 1A and 2A configurations, the compressor 1 may be used to produce a high temperature high pressure superheated fluid that condenses at a temperature of around 120 degrees. The condensed liquid passes into and is stored in liquid receiver 80. The refrigeration charge will be sufficient to hold liquid in the remaining parts of the refrigeration system for all piping and elements downstream of liquid receiver 80 until the expansion device 17. The liquid refrigerant passes through heat exchanger 11 and 51, each which reduces the sensible heat of the liquid refrigerant.

[0049] The inclusion of the water storage tank 36 may provide several benefits compared to prior art systems depending upon the deployment scenario. For example, the enhanced double hybrid heat pump system 50 including the tank 36 may:

1. eliminate the need for a contractor to size and purchase a separate buffer tank, therefore reducing project cost and complexity.

2. operate at different flow rates between the heat pump hot water condenser and the building load pump. Prior art heat pumps that do not use variable speed compressors have a manufacturer’s recommended minimum flow rate, typically in the range of 3 gallons per minute per ton. This flow rate may be challenging for many buildings, as a 5-ton system would require a minimum of 15 gallons per minute to flow through the system. This flowrate is very high for homes with existing boilers, which are accustomed to low flowrates and high differential temperatures. The result is that large pumps are used to meet the minimum flowrate, but the water is split between the structure 200 and a buffer tank.

3. soften the load fluctuations on the compressor without the use of variable frequency drive (“VFD”) components which add a huge amount of cost and complexity.

4. provide for thermal storage in the system, e.g., a 50-gallon tank may be used to store approximately 25,000 BTUs. 6. allow seamless integration between hydronic electric backup heat via resistance element 45 and the heat pump. Integrating the electric backup heat elements within the scope of the heat pump reduces the risk that an error by an installing contractor that would lead to the underuse of the heat pump.

7. allow for an indirect hot water heat exchanger to be used for domestic hot water preheat in the case where the tank is filled with hydronic water. In this example, the secondary benefit is that a double wall heat exchanger is not needed as there are 2 mediums between the refrigerant and the domestic water.

8. allow a large portion of the heat that would otherwise be wasted to be captured during heating and cooling operation. In cooling mode, this heat may be utilized for domestic hot water, such as by the indirect heat exchanger 41. If the heat pump is operating in cooling mode, the structure 200 may benefit from “free hot water” i.e., hot water as a byproduct of cooling, by recovering it in the storage tank 36 rather than a traditional air conditioning system which rejects the heat outside by a condenser unit. Also compared to most geothermal systems, which use a desuperheater for domestic hot water, this system may be capable of recovering 90% or greater of the heat normally rejected compared to only the superheat (typically less than 15% of the total). In both heating and cooling mode, the heat added by the compressor 1 first goes to this tank and may be used as needed with limited additional mechanical equipment (no additional mechanical equipment is needed to provide domestic hot water.)

9. allow a heat pump controller to sense the building load by the change in temperature without the sudden swings in the refrigerant pressure that would be caused by directly piping the hot water return to the heat pump.

[0050] The double hybrid system 50 of the present invention has several advantages for retrofitting residential structures that have existing hydronic heating infrastructure. These buildings may have hydronic heat emitters, typically baseboards, radiators or radiant floors, that were commonly sized at the time of installation based on higher temperature hot water; i.e., the 160-180 degrees that is a common supply temperature of a conventional boiler. These same heat emitters may have only a fraction of their original capacity when connected with supply water that is at a temperature typical of a heat pump system (110-120 degrees).

[0051] Various double hybrid heat pump system 50 embodiments may be configured to produce hot water for the hydronic systems and hot air for heating at the same time, from the same unit to overcome the heat deficiency created by using lower temperature water in the hydronic system. As previously described, by producing these two at the same time, the efficiency is greatly improved and the warm air may be used to supplement the heat emitters in a retrofitted building.

[0052] In addition, the double hybrid system 50 has further advantages in terms of overall system efficiency compared to prior art hot water only heat pumps. For example, a building that could be heated on a peak day with 120 degree hot water would require that the hot water only heat pump to deliver 1 0 degree hot water, whereas the double hybrid system 50 may be operated with a lower hot water temperature, for example to 100 or 110 degrees, and then supply heated forced air at the same time to provide the same total heating effect to the building. The skilled artisan will appreciate that a lower supply temperature of even a few degrees makes a large difference on heat pump efficiency. For example, a 10 degree reduction in supply temperature may increase the coefficient of performance by .25-5.

[0053] In addition, if the building’s heat emitters were severely limiting, the unit could cycle between producing only hot air and both hot air and hot water to ensure that the heat production of the unit is not limited by the capacity of the heat emitters in the building. The Split-Hybrid embodiments depicted in FIGS. 2A-2E may be used to supply heat in a variety of ways to suit the needs of a building. For instance, it may operate as a water to air heat pump and use a large ducted, refrigerant to air heat exchanger in an air handler to supply multiple rooms or apartments with heat, while also producing domestic hot water for the building. In other instances, a multi-family building may have 4 apartments, three that are heated by hot water and a fourth by forced air. The Split-Hybrid system embodiments may be used to simultaneously produce hot water for the 3 units while also producing hot air only for the 4 ^th .

[0054] The skilled artisan may employ other devices in the system 50. For example, a desuperheater may be employed between the compressor 1 and the condensing heat exchanger 3, to pre-cool the high-pressure, high temperature vapor prior to entering the condensing heat exchanger 3.

[0055] FIGS. 1A-2E depict embodiments of the system 50 configured as an enhanced performance heat pump, which include a refrigerant to refrigerant heat exchanger 51 between the refrigerant-second fluid heat exchanger 11 and the refrigerant to source heat exchanger 27. The inclusion of the refrigerant-to-refrigerant heat exchanger 51 serves to increase the subcooling of the refrigerant prior to entering the expansion device and the temperature of the gas returning to the compressor. [0056] The expansion device 17 may include an internal check valve in lieu of a separate check or bypass valve 54 to create a method of bypassing expansion device 17 in cooling mode. A mechanical bypass valve 53 is needed to allow refrigerant to bypass the expansion device 52 when it is desired to increase the heat output through heat exchanger 11. As depicted in FIG. IB and 1C, when refrigerant enters from 71 in heating mode, the bypass valve will be open, but closed in cooling mode. As depicted in Fig. 2A-2E, the system 50 may also include electronic valve 62 to bypass refrigerant partly or entirely from the heat exchanger 1 1 , which may be used to limit the amount of refrigerant flowing to heat exchanger 11 under reduced load conditions or to allow for hot water heat production only.

[0057] A control algorithm may be used to optimize the total system performance may include parameters such as the refrigerant evaporating and condensing temperatures, the relative interacting fluid temperatures, the temperatures of the refrigerant as it leaves the condenser and subcooler(s) and may control the expansion device 17 and 52, electronic control valves 62 and 53, the speeds of the pumps, 30 and 35 and blower(s) 12. The system 50 may include various sensors, e.g., pressure, temperature, etc. to provide data for control of the system 50, such as temperature sensors 90-93.

[0058] FIGS.S 2A-2E depict embodiments of the system 50, in which the refrigerantsecondary fluid heat exchanger 11 is embodied as a plurality of parallel refrigerant-secondary fluid heat exchanger modules 59, such as refrigerant- air heat exchangers, which may be referred to as blower modules. Each blower module contains a refrigerant to air heat exchanger, blower (not shown), electronic valve 53 and expansion device 52. A control device 60 may be provided in the flow paths to and from the blower modules 59 to regulate which blower modules 59 receive refrigerant. The blower modules 59 provide a split heat pump configuration that may be useful in various building, such as homes without a basement, multifamily buildings or buildings where ductwork cannot easily be run to essential locations. The blower modules 59 may be employed as high wall head modules, which have become popular in the air source heat pump industry. Each blower module 59 may include a separate expansion device 52, and bypass valve 53, so that each blower module 59 may be operated similar to the configuration of the heat exchanger 11 in the embodiments depicted in FIGS. 1A-1D.

[0059] In various embodiments, the compressor 1 produces high temperature, superheated gaseous refrigerant which is directed into the condensing heat exchanger 3. The condensing heat exchanger desuperheats and condenses the refrigerant and the heat energy is transferred to the water entering at port 33 of heat exchanger 3 and which returns to tank 36. When operating in heating mode, the refrigerant leaves condenser 3 and passes through the reversing device 15 and leaves from port 7 towards reversing device 61. If a system operator intends to produce high temperature hot water with supplemental hot air, the reversing device 61 directs refrigerant out of port 73 and into liquid receiver 80, where excess liquid refrigerant may be stored, as shown in FIGS. 2A. Liquid refrigerant leaves liquid receiver 80 and enters the control device 60, which directs refrigerant flow to the blower module(s) 59 that are calling for heat. The control device 60 may include multiple mechanical valves 63 to direct refrigerant as desired. Refrigerant enters the blower module 59 at port 56 and the refrigerant passes through the refrigerant-to-secondary fluid heat exchanger 11, which may be a refrigerant- air heat exchanger including a blower (not shown). Inside the module 59, the blower blows cool air across the heat exchanger 11 to further subcool the liquid refrigerant and provide heated air to the space associated with the blower module 59. Refrigerant passes around expansion device 52 through electronic bypass valve 53.

[0060] The refrigerant passes unrestricted by the expansion device 52 and returns to the control device 60 as a subcooled liquid refrigerant, where the multiple streams of refrigerant are recombined and is then directed towards heat exchanger 51 via reversing element 61 and enters through port 45. The liquid refrigerant is cooled by the relatively low pressure and temperature vapor (which is correspondingly superheated on the other side of heat exchanger 51) and leaves through port 18. The refrigerant may pass through a filter dryer 14 or may pass through a filter dryer 14 somewhere else in the refrigeration cycle, though generally they are placed near expansion devices.

[0061] The refrigerant continues to port 16, which may have a check valve 54 in parallel with the expansion device 17. The check valve 54 will not allow refrigerant to pass in this direction and so all refrigerant is directed through the expansion device 17. It should be noted that liquid refrigerant will stack or backfill starting from the active expansion device 17 and will backfill as a liquid until it reaches whichever heat exchanger is immediately upstream of the liquid receiver.

[0062] The high pressure liquid refrigerant passes through the expansion device 17 and exits as a low pressure saturated fluid containing both gas and liquid. When a large amount of subcooling is used, the amount of liquid being released into the low pressure/suction side of the refrigerant cycle increases. This saturated fluid is enters the refrigerant liquid receiver 19 and then is directed into the inlet port 26 of heat exchanger 27. Heat exchanger 27 acts as an evaporating heat exchanger and evaporates the liquid refrigerant into superheated gas. A traditional system may aim for 10 degrees Fahrenheit of superheat to ensure that no liquid is allowed into the compressor, since there are no components to further heat the refrigerant before it enters the compressor suction line. However, with an Enhanced Performance Double Hybrid, the expansion device 17 may aim for 1-4°F of superheat leaving the evaporating heat exchanger with the knowledge that the refrigerant will be further warmed by heat exchanger 51. By reducing the superheat, the refrigerant may operate at a ~6-9 ^c F higher saturation temperature and -12-16 psi higher suction pressure entering the evaporator, reducing the pressure difference (also called lift) for the compressor and decreasing the compression ratio of the compressor, both which are leading indicators of performance and compressor longevity. [0063] The superheated vapor exits the evaporator at port 25 and enters the reversing device 15 at port 8. The refrigerant is redirected out port 9 of a reversing device 15. The refrigerant enters heat exchanger 51 and is heated by subcooling the liquid refrigerant on the other side of the heat exchanger. The refrigerant enters the compressor suction line at port 29. [0064] FIG 2B depicts the same embodiments shown in FIG. 2A, but is configured to release a higher percentage of heat as forced air. In this mode, reversing valve 61 reverses the direction of the refrigerant through blower module 59. This in turn allows for the high pressure refrigerant to partially or fully condense in heat exchanger 3 and fully condense or subcool in heat exchangers 11 depending on the operation of those heat exchangers, such as the conditions and relative speeds of circulator pump 35 and blowers used in blower module 59. For instance, the system 50 may be configured to produce hot water and hot air heating at the same time. The refrigerant leaving heat exchanger 11 in blower module 59 will be fully condensed into a liquid and any excess will be stored in liquid receiver 80. The liquid refrigerant will leave the liquid receiver 80 and pass through heat exchanger 51, where it will be subcooled.

[0065] FIG 2C depicts the same embodiments shown in FIG. 2A, but in cooling mode. In this mode, refrigerant may be condensed exclusively at the condenser 3 by turning off the pump 30 or pump 30 may run slowly to allow the refrigerant to be subcooled in heat exchanger 27 after condensation in condenser 3. The refrigerant passes through liquid receiver 19 and extra refrigerant may be stored in this tank. The refrigerant is subcooled by heat exchanger 27 if the pump is operated slowly and by heat exchanger 51 prior to entering the control device 60. Subcooling the refrigerant prior to the control device 60 will increase the cooling capacity of the refrigerant inside the heat exchanger 11. In blower module 59, the expansion device 52 is active and refrigerant pressure is reduced on the outlet side of the expansion device. In addition, the superheated gas that enters port 48 of heat exchanger 51 will be further heated and create a larger discharge superheat at the outlet of the compressor 1 which may be used to create a higher temperature fluid from condenser 3.

[0066] One benefit of the enhanced performance embodiments, such as those depicted and described in FIGS 1C and 2D is in the production of higher temperature domestic hot water as a byproduct of cooling and with minimal drop in cooling efficiency. For example, the system 50 may be configured with pump 35 on and pump 30 off. The refrigerant to refrigerant heat exchanger 51 will absorb heat energy from the high pressure side of the refrigeration cycle, increasing cooling capacity and efficiency. The transferred heat will increase the discharge superheat from the compressor 1, which in turn will allow for the production of higher temperature hot water at port 34.

[0067] In FIGS. 1A-2E, different amounts of refrigerant may be required depending on the operation of the heat pump.

[0068] For instance, in FIGS. 1A and 2A, operation of pump 35 and condensing heat exchanger 3 may result in liquid accumulation behind the expansion device 17 at port 16 and back into the refrigerant- second fluid heat exchanger 11. The refrigerant charge may be sufficient to cause liquid collection in the liquid receiver 80 from heat exchanger 3. Adding a total volume of refrigerant that is greater than the volume of the refrigerant- secondary fluid heat exchanger 11 and all downstream piping is a useful method for forcing refrigerant subcooling.

[0069] Furthermore, as expected, the discharge refrigerant pressure depends on the temperatures and flowrates of the fluid moving through condenser 3. Without sufficient charge and a method of storing excess liquid refrigerant, the high pressure gas leaving the compressor 1 will condense at a temperature that is satisfactory for achieving volumetric balance in the system.

[0070] The refrigerant volume may be leveraged to force liquid through a series of progressively cooler heat exchangers, such depicted in FIGS. 1A-2E. In these embodiments, liquid refrigerant may accumulate from the inlet of the active expansion device, 17 or 52, back toward the condenser 3. [0071] For example, in FIG 1A embodiments, the refrigerant may have a pressure of 425 PSI (R410a) and may allow water passing through condenser 3 to be heated to 120°F or beyond. The liquid refrigerant leaving condenser 3 may be at a temperature of 120°F and may be cooled to 75°F when passing through refrigerant-secondary fluid heat exchanger 11 that warms air in a building by subcooling the liquid refrigerant. The subcooled refrigerant then passes through the subcooling refrigerant-to-refrigerant heat exchanger 51 that further cools the high pressure refrigerant to 40 ^c F. The resulting liquid that enters the expansion device 17 will be subcooled by approximately 80°F, which allows the low-pressure refrigerant on the low pressure side of the expansion device 17 to have a higher refrigerant quality (liquid to gas ratio) as it enters heat exchanger 27, which is acting as an evaporator. Higher refrigerant quality extracts more heat from the fluid on the other side of the evaporator, typically water or air.

[0072] FIG 1C allows this same subcooling process but in cooling mode. Since it is unlikely that the heating and cooling modes will require the same amount of refrigerant, excess refrigerant will accumulate in the liquid receiver 19. Since the liquid receiver 19 is downstream of both the condenser 3 and heat exchanger 27, which serves as a condenser in the heating mode, either of these heat exchangers may be used as a primary condenser or they may be used in series where conditions are suitable and heat exchanger 27 may subcool the refrigerant leaving condenser 3.

[0073] FIGS. 3A & 3B show exemplary refrigeration heating and cooling cycles, respectively, in terms of pressure versus enthalpy for the prior art and present invention. The solid black line is a standard prior art water to water heat pump system and dashed-dot purple lines show the refrigeration cycle with subcooling and condensing heat exchangers. The greater width of the refrigeration cycle in FIG. 3A and 3B is due to the subcooling heat exchanger (HE), which provides additional heat transfer from the refrigerant after leaving condensing heat exchanger 3. The net effect created is a refrigeration cycle that has a larger refrigeration effect with lower compressor work, resulting in higher efficiency. One of ordinary skill in the art will appreciate that the refrigerant leaving the condensing heat exchanger 3 which is passed to the cooling heat exchanger, and the refrigerant exiting the expansion device 17 may be 100% liquid or a two-phase mixture as desired. In exemplary cycle 3B, the condensing heat exchanger 3 absorbs the latent heat from the high-pressure fluid and then subcooling heat exchanger 27 further cools the refrigerant within a few degrees of the ground loop temperature. [0074] FIG 3A shows an exemplary refrigeration heating cycle with the addition of heat exchanger 51 to refrigerant to refrigerant subcooling. As one familiar in the art will understand, the increasing the width of the refrigeration cycle increases the amount of heat that can be extracted. Decreasing the refrigeration lift decreases the compressor power input, both of which increase the system efficiency. The addition of refrigerant-to-refrigerant heat exchanger 51 allows for both of these favorable characteristics compared with a standard refrigeration cycle. The heat absorbed from the second subcooling heat exchanger 51 allows for greater heat absorption from the refrigerant on the evaporating heat exchanger 27. This heat is transferred to the gaseous refrigerant leaving the evaporator, increasing the superheat prior to the gas entering the compressor 1. Higher superheat entering the compressor 1 will cause the refrigerant to have a higher discharge superheat, which will increase the temperature of the fluid in contact with the first condenser. Since the refrigerant exiting the principal evaporator is being superheated in both heating and cooling modes, the refrigerant pressure in the evaporator may be increased and the expansion device may be allowed to target a lower-than- normal superheat, knowing that the refrigerant will be further superheated by heat exchanger 51 prior to entering the compressor.

[0075] For instance, a common superheat setpoint for residential heat pumps is 10°F. Since the refrigerant will be further superheated, a manufacturer may choose to reduce this to 2-4°F, accepting that any small droplets of refrigerant that could pass through will be later heated by the refrigerant-to-refrigerant heat exchanger, and that the final superheat entering the compressor could be in excess of 20°F.

[0076] To further illustrate the value of this subcooling, the following are the enthalpies of the refrigerant with a traditional heat pump vs an enhanced Double Hybrid heat pump:

Traditional heat pump heating water to 120°F

Enthalpy of Superheated refrigerant heated to 185 °F = 480 kJ/kg Enthalpy of condensed liquid refrigerant at 120°F = 284 kJ/kg Enthalpy of condensed liquid refrigerant at 115°F, (typical subcooling) = 280 kJ/kg Total heat released to system: 480 kJ/kg - 284 kJ/kg = 200 kJ/kg

Enhanced Double Hybrid heat pump heating water to 120 ^Q F Enthalpy of Superheated refrigerant heated to 210°F = 499 kJ/kg Enthalpy of condensed liquid refrigerant at 120°F = 284 kJ/kg Enthalpy of condensed liquid refrigerant at 80°F, (HX 11 subcooling) = 244 kJ/kg Enthalpy of condensed liquid refrigerant at 56°F, (HX 51 subcooling) = 225 kJ/kg* Total heat released to system: 499 kJ/kg - 244 kJ/kg = 255 kJ/kg

*These BTUs are transferred to the superheated vapor prior to entering the compressor 1 and are accounted for in the higher superheated gas enthalpy.

[0077] It can be seen that 55 kJ/kg of additional heat is released with an enhanced double hybrid and an overall increase in heat released of 27.5%. In addition, the compressor lift is reduced since the evaporator may operate with a higher pressure further improving performance. Though the effect is small, this example also illustrates how higher temperature water can be produced by using the higher temperature superheated vapor. With the same condensing temperatures, the Enhanced Double Hybrid may produce 122°F or 123 °Fhot water whereas the traditional system may only make 120°F.

[0078] FIGS. 4A-4D show a prior art embodiment (4 A) and embodiments of the present invention (4B-4D) depicted in various installation scenarios. FIG. 4A shows a typical prior art retrofit of a boiler based heating system with a conventional heat pump. FIGS 4B & 4C depict various embodiments of the system 100 with a DHHP 50 installed in a structure 200 in the heating (4B) and cooling (4C) modes, respectively. FIG. 4B shows the flow of heat in the structure 200 in the heating mode and provides exemplary use cases. As previously discussed, and shown in FIG. 4B, the system 100 with a DHHP 50 provides hot water to the boiler, or in lieu of the boiler, to support hot water heating and forced hot air for additional heating. While FIG. 4B shows ductwork and forced hot air being provided to the first floor only, one of ordinary skill will appreciate the ductwork may be provided to the higher floors as desired to meet design and budget objectives.

[0079] FIG. 4C shows the flow of cool air in the structure 200 in the cooling mode and the provision of hot water for domestic hot water uses. These embodiments are particularly efficient as the heat is removed from the air in the structure 200 may be used to heat water for domestic hot water use.

[0080] While the present invention was described in various embodiments as being used in combination with a boiler, unlike prior art units, the DHHP 50 of the present invention may be used to replace an existing boiler. The integration of the refrigerant subcooling function may enable substantially higher coefficients of performance (COP), such as 3.5-4.5 in regular heating mode and up to 12-18 in cooling mode (accounting for the hot water benefit) compared of COP of between 2.8-3.2 in heating mode and 4.5-6 in cooling mode that is typical of water to water heat pumps with a desuperheater. In addition, system 100 employing the DHHP system 50 of the present invention are simpler to install resulting in a lower risk of job failure or recall for the contractor.

[0081] One of skill in the art will appreciate that the system 50 may be implemented with fixed or variable speed pumps and blowers 12 to provide flexibility in the operation and control. For example, the system 50 may employ variable speed blowers 12, which typically have a lower parasitic electric load than multiple small single speed fans that may be used by hydronic air handlers and more flexibility than one fixed speed fan. Likewise, one more pumps 30 and 35 in the system 50 may be variable speed.

[0082] The foregoing disclosure provides examples, illustrations and descriptions of the present invention, but is not intended to be exhaustive or to limit the implementations to the precise form disclosed. Modifications and variations are possible in light of the above disclosure or may be acquired from practice of the implementations. These and other variations and modifications of the present invention are possible and contemplated, and it is intended that the foregoing specification and the following claims cover such modifications and variations.

[0083] As used herein, the term component is intended to be broadly construed as hardware, firmware, and/or a combination of hardware and software. It will be apparent that systems and/or methods, described herein, may be implemented in different forms of hardware, firmware, or a combination of hardware and software. The actual specialized control hardware or software code used to implement these systems and/or methods is not limiting of the implementations. Thus, the operation and behavior of the systems and/or methods were described herein without reference to specific software code— it being understood that software and hardware may be designed to implement the systems and/or methods based on the description herein.

[0084] Some implementations are described herein in connection with thresholds. As used herein, satisfying a threshold may refer to a value being greater than the threshold, more than the threshold, higher than the threshold, greater than or equal to the threshold, less than the threshold, fewer than the threshold, lower than the threshold, less than or equal to the threshold, equal to the threshold, etc.

[0085] Even though particular combinations of features are recited in the claims and/or disclosed in the specification, these combinations are not intended to limit the disclosure of possible implementations. In fact, many of these features may be combined in ways not specifically recited in the claims and/or disclosed in the specification. Although each dependent claim listed below may directly depend on only one claim, the disclosure of possible implementations includes each dependent claim in combination with every other claim in the claim set.

[0086] No element, act, or instruction used herein should be construed as critical or essential unless explicitly described as such. Also, as used herein, the articles "a" and "an" and the term "set" is intended to include one or more items and may be used interchangeably with “one or more”. Where only one item is intended, the term "one" or similar language is used. Also, as used herein, the terms "has," "have," "having," or the like are intended to be open- ended terms. Further, the phrase "based on" is intended to mean "based, at least in part, on" unless explicitly stated otherwise.