Patent Application
20210033315
The disclosure relates to heat pumps. More particularly, the disclosure relates to heat pumps for use with low pressure refrigerants.
Some vapor compression systems are configured to alternatively operate in a heating mode and a cooling mode. For example, in a residential heat pump situation, a cooling mode involves a compressor receiving refrigerant from an indoor heat exchanger and compressing that refrigerant. The compressed refrigerant passes through an outdoor heat exchanger where it rejects heat and is condensed. The condensed refrigerant passes through an expansion device and further cools. The expanded/cooled refrigerant absorbs heat in the indoor heat exchanger.
In a heating mode, the compressor receives and compresses refrigerant from the outdoor heat exchanger. The compressed refrigerant rejects heat in the indoor heat exchanger before passing rough the expansion device and outdoor heat exchanger.
Commercial chiller units have similar mode switching, but have one or both of the heat exchangers as refrigerant-water heat exchangers.
To facilitate switchover between cooling and heating modes, the system will typically include a four-way switching/reversing valve or an alternative combination of two-way and/or three-way valves.
Efforts have been made to develop systems using low global warming potential (GWP) refrigerants. One example of a low-GWP refrigerant is R1233zd(E) (hereafter simply “R1233zd”). Whereas, R410A has a direct GWP of 2088, R1233zd has a direct GWP of less than 1.0. R1233zd also has a higher cycle efficiency than R410A (e.g., by about 10% to 15%) due to lower discharge temperatures and lower expansion losses. Nevertheless, R1233zd suffers from being a low pressure refrigerant. A low pressure refrigerant is defined by the United States Environmental Protection Agency (EPA) as having a saturation pressure less than 45 psia (310 kPa) (R1233zd has a saturation pressure of 31 psia (214 kPa)) at 104° F. (40° C.). Low pressure refrigeration systems typically operate at evaporator pressures (thus compressor suction pressures) less than atmospheric pressure or the ambient pressure which might slightly differ from 1 ATM due to weather, altitude, etc.
R410A (saturation pressure 352 psia (2.43 MPa) at 104° F. (40° C.)) is a high-pressure refrigerant (saturation pressure 170 psia (1.17 MPa) to 355 psia (2.45 MPa) at 104° F. (40° C.)). R134a (saturation pressure 147 psia (1.01 MPa) at 104° F. (40° C.)) is a medium pressure refrigerant (saturation pressure 45 psia (310 kPa) to 170 psia (1.17 MPa) at 104° F. (40° C.)).
R1233zd also has benefits of being non-flammable and nontoxic (rating A1 under ASHRAE Standard 34-2007; with “A” indicating non-toxic and “1” indicating non-flammability).
One aspect of the disclosure involves a vapor compression system comprising: a first compressor; a second compressor; a first heat exchanger; a second heat exchanger; and one or more expansion devices. Means are provided for switching the system between operation in a first mode and a second mode. In the first mode: the first compressor compresses refrigerant; the compressed refrigerant is cooled in the first heat exchanger; the cooled refrigerant is expanded in at least one of the one or more expansion devices; the expanded refrigerant absorbs heat in the second heat exchanger and returns to the first compressor; and the second compressor is offline. In the second mode: the second compressor compresses refrigerant; the compressed refrigerant is cooled in the second heat exchanger; the cooled refrigerant is expanded in at least one of the one or more expansion devices; the expanded refrigerant absorbs heat in the first heat exchanger and returns to the second compressor; and the first compressor is offline.
In one or more embodiments of any of the foregoing embodiments, the first compressor and the second compressor share an inverter.
In one or more embodiments of any of the foregoing embodiments, the first compressor and the second compressor share a motor.
In one or more embodiments of any of the foregoing embodiments, the first compressor and the second compressor are respectively coupled to the motor by a first clutch and a second clutch.
In one or more embodiments of any of the foregoing embodiments, the first heat exchanger is an outdoor heat exchanger and the second heat exchanger is an indoor heat exchanger.
In one or more embodiments of any of the foregoing embodiments, the first heat exchanger is a refrigerant-air heat exchanger.
In one or more embodiments of any of the foregoing embodiments, the second heat exchanger is a refrigerant-liquid heat exchanger.
In one or more embodiments of any of the foregoing embodiments, the second heat exchanger is a refrigerant-air heat exchanger.
In one or more embodiments of any of the foregoing embodiments, the second compressor has a pressure ratio at least 1.25 times a pressure ratio of the first compressor.
In one or more embodiments of any of the foregoing embodiments, the first compressor is a scroll compressor; and the second compressor is a screw compressor or a centrifugal compressor.
In one or more embodiments of any of the foregoing embodiments, the first compressor and the second compressor are both screw compressors; or the first compressor and the second compressor are both centrifugal compressors.
In one or more embodiments of any of the foregoing embodiments, the one or more expansion devices comprise: a first expansion device not passing refrigerant in the second mode; and a second expansion device not passing refrigerant in the first mode.
In one or more embodiments of any of the foregoing embodiments, the system contains a low pressure refrigerant.
In one or more embodiments of any of the foregoing embodiments, the system is a chiller.
In one or more embodiments of any of the foregoing embodiments, a method for using the system comprises: running the system in the first mode; and running the system in the second mode.
In one or more embodiments of any of the foregoing embodiments, the first mode is a cooling mode and the second mode is a heating mode.
In one or more embodiments of any of the foregoing embodiments, in at least one of the first mode and the second mode, a compressor suction pressure is less than an ambient pressure.
Another aspect of the disclosure involves a method for operating a vapor compression system. The vapor compression system comprises: a first compressor; a second compressor; a first heat exchanger; a second heat exchanger; and one or more expansion devices. The method comprises running the system in a first mode and running the system in a second mode. In the first mode: the first compressor compresses refrigerant; the compressed refrigerant is cooled in the first heat exchanger; the cooled refrigerant is expanded in at least one of the one or more expansion devices; the expanded refrigerant absorbs heat in the second heat exchanger and returns to the first compressor; and the second compressor is offline. In the second mode: the second compressor compresses refrigerant; the compressed refrigerant is cooled in the second heat exchanger; the cooled refrigerant is expanded in at least one of the one or more expansion devices; the expanded refrigerant absorbs heat in the first heat exchanger and returns to the second compressor; and the first compressor is offline.
In one or more embodiments of any of the foregoing embodiments, the first mode is a cooling mode and the second mode is a heating mode.
In one or more embodiments of any of the foregoing embodiments, in at least one of the first mode and the second mode, a compressor suction pressure is less than an ambient pressure.
In one or more embodiments of any of the foregoing embodiments, switching between the first mode and the second mode comprises switching a single inverter between powering the first compressor and the second compressor; and said switching between the first mode and the second mode does not involve use of a four-way reversing valve.
The details of one or more embodiments are set forth in the accompanying drawings and the description below. Other features, objects, and advantages will be apparent from the description and drawings, and from the claims.
Like reference numbers and designations in the various drawings indicate like elements.
Low pressure refrigerants pose problems for use in systems having four-way reversing valves due to large valve size and associated losses. As is discussed below, a set of compressors and valves are used to accomplish the function of a traditional four-way reversing valve and facilitate heat pump operation. In typical state-of-the-art reversible heat pumps multiple compressors are already used to meet capacity and/or operational envelope. With low pressure refrigerants, current technology would require multiple compressors (either in parallel or in series) in order to meet the design envelope associated with the dramatic change of thermodynamic properties. Low pressure refrigerants require compression at higher pressure ratios than do high pressure refrigerants. Lower pressure refrigerants may require a greater spread between the lowest cooling mode pressure ratio and the highest heating mode pressure ratio than do high pressure refrigerants.
The exemplary heat exchangers 40 and 46 are refrigerant-water heat exchangers wherein the refrigerant transfers heat with a liquid such as water, brine, or glycol. The term “refrigerant-water heat exchanger” is often used in the art when the heat transfer fluid is a liquid other than pure water. Thus, the first heat exchanger 40 has a pair of liquid ports 56, 58 and the second heat exchanger 46 has a pair of liquid ports 60, 62 for respectively passing liquid flows. The ports respectively pass flows 64 and 66 through the two heat exchangers along heat transfer legs 68, 70 in heat exchange relationship with the refrigerant flows through the respective heat exchanger.
In the cooling mode, refrigerant is compressed by the first compressor 22A and passes out the discharge port 26A along the discharge line 30, passing through the junction 36 to the port 42. Within the heat exchanger 40, the refrigerant rejects heat to the liquid flow 64 and exits the port 44. Typically, the refrigerant will condense in the heat exchanger 40. Thus, the heat exchanger 40 serves as a condenser so that vapor refrigerant enters the port 42 but liquid refrigerant exits the port 44.
The refrigerant passes from the port 44 through the refrigerant line to the expansion device 52 where it expands and its temperature drops. The refrigerant then passes into the second heat exchanger 46 via the port 48. In the second heat exchanger, the refrigerant absorbs heat from the liquid flow 66 and exits the port 50. In the second heat exchanger 46, the refrigerant evaporates so that the second heat exchanger functions as an evaporator. The refrigerant then passes through the junction 34 to the suction line 28A returning to the first compressor 22A to complete the compression cycle.
In the exemplary cooling mode, the liquid flow 66 may pass along a cooling loop to liquid-air heat exchangers throughout a building. The liquid 64 may pass to liquid-air heat exchangers outside to reject heat to an external environment.
In alternative embodiments, the first heat exchanger 40 may be a refrigerant-air heat exchanger directly discharging heat to external air (e.g., a fan-forced outdoor air flow driven by an electric fan (not shown)) in the cooling mode. The second heat exchanger 46 may still be a refrigerant-liquid heat exchanger as in a chiller (e.g., a heat pump chiller). Or the second heat exchanger may also be a refrigerant-air heat exchanger such as in a forced-air residential or commercial heat pump. An exemplary such commercial heat pump is a rooftop unit having a pull-through fan or a push-through fan to drive the outdoor/external air flow.
Similarly, in further variations, the liquid flow 66 may not directly absorb heat from the air within the building but may in turn serve as the heat absorbing fluid for individual further vapor compression systems operating with refrigerant-liquid condensers and refrigerant-air evaporators.
In the heating mode, however, the first compressor is off and the second compressor is on and the flow proceeds in a reverse direction through the ports of the two heat exchangers and the expansion device 52. Thus, the heat exchanger 46 serves as a heat rejection heat exchanger rejecting heat to the liquid 66 which in turn rejects heat in the liquid-air heat exchangers within the building. Similarly, the heat exchanger 40 serves as an evaporator absorbing heat.
The controller 100 may receive user inputs from an input device (e.g., switches, keyboard, or the like) and sensors (not shown, e.g., pressure sensors and temperature sensors at various system locations). The controller may be coupled to the sensors and controllable system components (e.g., valves, the bearings, the compressor motor, vane actuators, and the like) via control lines (e.g., hardwired or wireless communication paths). The controller may include one or more: processors; memory (e.g., for storing program information for execution by the processor to perform the operational methods and for storing data used or generated by the program(s)); and hardware interface devices (e.g., ports) for interfacing with input/output devices and controllable system components.
The system may be made using otherwise conventional or yet-developed materials and techniques.
The two compressors may be optimized for respective cooling and heating mode operation. Asymmetries in the compressors are particularly relevant where a low pressure refrigerant is used (e.g., R1233zd or other low pressure refrigerant). A desired pressure ratio for the heating mode will be significantly higher than that for the cooling mode. Thus, for example, the pressure ratio of the second compressor 22B may be an exemplary 1.25 to 10.0 times, more particularly 2.0 to 10.0 times, that of the first compressor 22A (or at least 1.25 times or at least 2.0 times).
The exemplary compressors are each variable-capacity compressors. For such variable capacity compressors, the relative pressure ratios may be measured at the maximum pressure ratio of each compressor. Each compressor may have its own electric motor and inverter (not shown) coupled to external power. The compressors may be controlled by a controller 100. Although the compressors may be of the same general type as each other (e.g., centrifugal), two different types may be used. The cooling mode compressor 22A could be scroll or screw or a low- or medium-lift centrifugal compressor. The heating mode compressor 22B may be a screw or centrifugal compressor but not likely a scroll because high pressure ratios required at extreme heating conditions may require excessive scroll speeds in view of the large scroll size needed for low pressure refrigerants.
A particular example of a combination for use with medium pressure refrigerants (e.g., R134a and the like) is a scroll compressor for the first compressor and screw compressor or centrifugal compressor for the second compressor. With low pressure refrigerants, both are likely to be screw or centrifugal. There may be one of each type or both may be the same type.
There may be further variations on the system 320 as the systems 120 and 220 are on the system 20.
As noted above, whereas typical prior art heat pumps utilize four-way switching/reversing valves, the illustrated embodiments may avoid such valves. This presents a particularly significant advantage when working with low pressure refrigerants. When such a low pressure refrigerant is used, the four-way valve would have to be very large (thus expensive) and would impose significant energy losses. By using the two different compressors along with merely on-off valves (controllable valves (e.g., solenoid valves) and/or check valves), this expense and loss of efficiency may be avoided.
Additionally, in various implementations relative to compressors having switch over valves, the ability to tailor one compressor to each mode may further improve efficiency.
Although a single respective compressor is shown for each mode, one or both of the modes may have multiple compressors in series, parallel, or otherwise.
The use of “first”, “second”, and the like in the description and following claims is for differentiation within the claim only and does not necessarily indicate relative or absolute importance or temporal order. Similarly, the identification in a claim of one element as “first” (or the like) does not preclude such “first” element from identifying an element that is referred to as “second” (or the like) in another claim or in the description.
Where a measure is given in English units followed by a parenthetical containing SI or other units, the parenthetical's units are a conversion and should not imply a degree of precision not found in the English units.
One or more embodiments have been described. Nevertheless, it will be understood that various modifications may be made. For example, when applied to an existing basic system, details of such configuration or its associated use may influence details of particular implementations. Accordingly, other embodiments are within the scope of the following claims.
Benefit is claimed of U.S. Patent Application No. 62/658,493, filed Apr. 16, 2018, and entitled “Dual Compressor Heat Pump”, the disclosure of which is incorporated by reference herein in its entirety as if set forth at length.
| Filing Document | Filing Date | Country | Kind |
|---|---|---|---|
| PCT/US2019/022193 | 3/14/2019 | WO | 00 |
| Number | Date | Country | |
|---|---|---|---|
| 62658493 | Apr 2018 | US |
