U.S. patent application number 13/968028 was filed with the patent office on 2015-02-19 for partitioned evaporator for a reversible heat pump system operating in the heating mode.
The applicant listed for this patent is Heat Pump Technologies, LLC. Invention is credited to Luther D. Albertson.
Application Number | 20150047385 13/968028 |
Document ID | / |
Family ID | 52465831 |
Filed Date | 2015-02-19 |
United States Patent
Application |
20150047385 |
Kind Code |
A1 |
Albertson; Luther D. |
February 19, 2015 |
PARTITIONED EVAPORATOR FOR A REVERSIBLE HEAT PUMP SYSTEM OPERATING
IN THE HEATING MODE
Abstract
Disclosed are embodiments of a partitioned evaporator for a
reversible heat pump system operating in the heating mode.
Embodiments include a partitioned evaporator having a first part
and a second part, the first part exchanging heat between a working
fluid upstream from a reversible metering device, and the working
fluid downstream from the reversible metering device. Heat energy
from the upstream fluid is recovered and introduced into the first
and second parts aiding in the evaporation of the working fluid
rather than being rejected and unused.
Inventors: |
Albertson; Luther D.;
(Sellersburg, IN) |
|
Applicant: |
Name |
City |
State |
Country |
Type |
Heat Pump Technologies, LLC |
Franklin |
IN |
US |
|
|
Family ID: |
52465831 |
Appl. No.: |
13/968028 |
Filed: |
August 15, 2013 |
Current U.S.
Class: |
62/324.6 |
Current CPC
Class: |
Y02P 80/10 20151101;
F25B 40/00 20130101; F25B 13/00 20130101; Y02P 80/156 20151101;
F25B 39/028 20130101; F25B 2313/02741 20130101 |
Class at
Publication: |
62/324.6 |
International
Class: |
F25B 13/00 20060101
F25B013/00; F25B 30/02 20060101 F25B030/02 |
Claims
1. A reversible heat pump system having 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,
comprising: a compressor coupled to a downstream condenser; and a
partitioned evaporator having a first part exchanging heat between
a working fluid from the condenser at a first temperature and the
working fluid from a metering device at a second temperature, the
metering device receiving 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; and a second part downstream from the first part and
upstream from the compressor, the second part exchanging heat with
an external medium; wherein 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 a heating mode to heat an enclosed space.
2. The heat pump system of claim 1, wherein the second delta is
less than about 5 times larger than a first delta.
3. The heat pump system of claim 1, wherein the second delta is
less than about 3 times larger than a first delta.
4. The heat pump system of claim 1, wherein the second delta is
less than about 2 times larger than a first delta.
5. The heat pump system of claim 1, wherein the second part is
downstream from the metering device and receives working fluid
directly from the metering device and from the first part.
6. The heat pump system of claim 1, wherein the first part
exchanges heat between the working fluid upstream and downstream of
the metering device using any one of the following: a tube-in-tube
heat exchanger, a coaxial coil heat exchanger, a plate heat
exchanger, or a shell-and-tube heat exchanger.
7. The heat pump system of claim 1, wherein the external medium is
any one of the following: ambient air, a liquid, or earth.
8. A reversible heat pump system having 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, comprising: a condenser
downstream from a compressor; and a multi-stage evaporator
downstream from the condenser and upstream from the compressor
having; 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 receiving the working fluid from the first stage at
a downstream temperature, the first stage heat exchanger exchanging
heat so that the downstream temperature of the working fluid is at
least about 10 percent less than the upstream temperature during
normal operations; and 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; wherein 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 a
heating mode to heat an enclosed space.
9. The heat pump system of claim 8, wherein the downstream
temperature of the working fluid is at least about 15 percent less
than the upstream temperature during normal operations.
10. The heat pump system of claim 8, wherein the downstream
temperature of the working fluid is at least about 20 percent less
than the upstream temperature during normal operations.
11. The heat pump system of claim 8, wherein the first heat
exchanger exchanges heat between the working fluid upstream and
downstream of the metering device using any one of the following: a
tube-in-tube heat exchanger, a coaxial coil heat exchanger, a plate
heat exchanger, or a shell-and-tube heat exchanger.
12. The heat pump system of claim 8, wherein the external medium is
any one of the following: ambient air, a liquid, or earth.
13. The heat pump system of claim 8, wherein the second heat
exchanger is downstream from the metering device and receives
working fluid directly from the metering device and from the first
heat exchanger.
14. A reversible heat pump system 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, comprising: a two-stage evaporator having 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 exchanging 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, the first stage exchanging at least about 10 percent
of heat exchanged by the second stage so that the downstream
temperature is less than the upstream temperature, the second stage
exchanging heat between the working fluid downstream from the first
stage and an external medium; and a compressor upstream from the
condenser and downstream from the second stage of the two-stage
evaporator, wherein 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 a heating mode
to heat an enclosed space.
15. The heat pump system of claim 14, wherein the first stage
exchanges at least about 15 percent of heat exchanged by the second
stage.
16. The heat pump system of claim 14, wherein the first stage
exchanges at least about 20 percent of heat exchanged by the second
stage.
17. The heat pump system of claim 14, wherein the downstream
temperature of the working fluid is at least about 10 percent less
than the upstream temperature during normal operations.
18. The heat pump system of claim 14, wherein the downstream
temperature of the working fluid is at least about 20 percent less
than the upstream temperature during normal operations.
19. The heat pump system of claim 14, wherein the first stage
exchanges heat between the working fluid from the condenser and the
working fluid downstream from the metering device using any one of
the following: a tube-in-tube heat exchanger, a coaxial coil heat
exchanger, a plate heat exchanger, or a shell-and-tube heat
exchanger.
20. The heat pump system of claim 14, wherein the external medium
is any one of the following: ambient air, a liquid, or earth.
Description
BACKGROUND
[0001] 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.
[0002] 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.
[0003] 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.
[0004] 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.
[0005] 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.
[0006] 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.
[0007] 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.
[0008] 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.
SUMMARY
[0009] 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.
[0010] 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.
[0011] 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.
[0012] 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.
[0013] 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.
[0014] Further forms, objects, features, aspects, benefits,
advantages, and embodiments will become apparent from the detailed
description and drawings provided herewith.
BRIEF DESCRIPTION OF THE DRAWINGS
[0015] FIG. 1A is a schematic view of a reversible heat pump system
having a partitioned evaporator operating in the heating mode.
[0016] FIG. 1B is a schematic view of the reversible heat pump
system of FIG. 1A operating in the cooling mode.
[0017] FIG. 2 is a diagrammatic view of one embodiment of a
partitioned evaporator like the partitioned evaporator shown in
FIGS. 1A and 1B.
[0018] FIG. 3 is a diagrammatic view of another embodiment of a
partitioned evaporator similar to the partitioned evaporator shown
in FIG. 2.
[0019] FIG. 4 is a perspective view of yet the embodiment of a
partitioned evaporator shown in FIG. 3.
[0020] FIG. 5A is a diagrammatic view of another embodiment of a
partitioned evaporator like the partitioned evaporator shown in
FIGS. 1A and 1B.
[0021] FIG. 5B is a diagrammatic view of another embodiment of a
partitioned evaporator like partitioned evaporator shown in FIG.
5A.
[0022] FIG. 6 is diagrammatic view of another embodiment of a
partitioned evaporator like the partitioned evaporator shown in
FIGS. 1A and 1B.
DETAILED DESCRIPTION
[0023] 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.
[0024] 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.
[0025] 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.
[0026] 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.
[0027] 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.TM. Suva.RTM. 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).
[0028] 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).
[0029] 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.TM. Suva.RTM. 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.
[0030] 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.
[0031] 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.
[0032] 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.
[0033] 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.
[0034] 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).
[0035] 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.
[0036] 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.
[0037] 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.
[0038] 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.
[0039] 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.
[0040] 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.
[0041] 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.
[0042] 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.
[0043] 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 FIG. 1, an element identified by a "200"
series reference numeral will first appear in FIG. 2, and so on.
With reference to the Specification, Abstract, and Claims sections
herein, it should be noted that the singular forms "a", "an",
"the", and the like include plural referents unless expressly
discussed otherwise. As an illustration, references to "a device"
or "the device" include one or more of such devices and equivalents
thereof.
[0044] FIGS. 1A and 1B illustrate in a schematic form the major
components of the disclosed system operating in the heating (FIG.
1A) and cooling mode (FIG. 1B). In FIG. 1A reversible heat pump
system 100 operates in the heating mode using a compressor 107,
reversing valve 109 downstream from compressor 107 for reversing
the flow of a working fluid or refrigerant 113, and an indoor heat
transfer unit 111 downstream from reversing valve 109. In the
heating mode, indoor heat transfer unit 111 operates as a condenser
rejecting heat 116 into an enclosed space 114 to raise the
temperature inside. A reversible metering device 110 is positioned
downstream from indoor heat transfer unit 111 and upstream from an
outdoor heat transfer unit 102.
[0045] 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.
[0046] 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 FIG. 1A (or FIG. 1B) for clarity.
[0047] 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.
[0048] 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.
[0049] 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.
[0050] 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.
[0051] 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 FIG.
1B, reversible metering device 106 does not cause any substantial
reduction in flow or working fluid pressure while reversible
metering device 110 does.
[0052] Considering the cooling mode illustrated in FIG. 1B,
reversing valve 109, as well as reversible metering devices 110 and
106 are configured to reverse the flow of working fluid 113 through
the closed refrigeration circuit to cool the enclosed space 114
rather than heat it. Heat 116 is rejected into an external medium
(for example, ambient air) from second part 105 causing the working
fluid vapor 113 to condense into a warm fluid outside enclosed
space 114. The warm liquid working fluid then passes through
reversible metering device 106 which is now reconfigured to avoid a
substantial pressure reduction of working fluid 113 as in the
heating mode. Working fluid 113 then passes through first part 103
as well, either directly from second part 105 or through reversible
metering device 106. Because metering device 106 no longer operates
to reduce the pressure between the first part and the second part
of outdoor heat transfer unit 102, first part 103 causes no heat
exchange between incoming and outgoing working fluid as in the
heating mode because both incoming and outgoing working fluid 113
passes through the one or more heat exchange paths in first part
103 carrying relatively warm condensed liquid at substantially the
same temperature. Thus in the cooling mode, outdoor heat transfer
unit 102 operates as a condenser and indoor heat transfer unit 111
operates as an evaporator.
[0053] No specific implementation of any of the dimensions or
components should be inferred from FIGS. 1A and 1B. For example,
compressor 107 may be any device useful for increasing the pressure
of working fluid 113, such as, for example, by reducing its volume.
Such devices include, but are not limited to, various types of
rotary compressors such as lobe compressors, screw compressors,
liquid ring compressors, scroll compressors, or vane compressors.
Other types of compressors and include reciprocating compressors
such as diaphragm compressors, double acting and single acting
reciprocating compressors such as piston or swash plate
compressors. Other types of compressors include centrifugal axial
compressors. These are but a few nonlimiting possibilities of
various embodiments of compressor 107.
[0054] 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
FIGS. 1A and 1B. Examples include but are not limited to,
Chlorodifluoromethane (CHClF2) carrying the ASHRAE designation
R-22, Tetrafluoroethane (C2H2F4) referred to by ASHRAE as R-134,
the above mentioned R-410A, or a mixture of R-22 and
Chloropentafluoroethane (C2F5Cl) referred to by ASHRAE as R-115,
the mixture including about 48.8 percent R-22 and about 51.2
percent R-115 and referred to by ASHRAE as R-502. These nonlimiting
examples along with any other suitable chemical compositions having
properties advantageous to the particular heat pump system are
envisioned for working fluid 113.
[0055] 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.
[0056] 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.
[0057] 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.
[0058] Reversible metering devices 110 and 106 illustrated in FIGS.
1A and 1B include any device or system for controlling the amount
of refrigerant flow into the indoor or outdoor heat transfer unit
102 or 111 respectively depending on which unit 102 or 111 is
operating as an evaporator. One embodiment of such a device is a
thermal expansion valve, although other devices with similar
properties and behavior are envisioned as well. Metering or flow
control of working fluid 113 may be accomplished by various means
such as using a temperature sensor coupled to the metering device.
Examples of such temperature sensors include a sensing bulb
containing a fluid similar to working fluid 113, or an electronic
sensing device, or other suitable apparatus for sensing temperature
of working fluid 113. The temperature sensor communicates the
temperature of the working fluid leaving the evaporator causing
metering devices 110 and 106 to open based on the temperature of
vaporized working fluid exiting the evaporating chamber. Metering
devices 110 and 106 may also include a bypass or check valve which
channels working fluid 113 around the metering components within
reversible metering devices 110 and 106 as the fluid flows away
from the condenser (111 in the heating mode, 102 in the cooling
mode).
[0059] 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.
[0060] Illustrated in FIGS. 2 and 3 are embodiments of a
partitioned, multi-stage, or two-stage evaporator 200 having a
first part (first stage) 203 and second part (second stage 205).
Multistage evaporator 200 is configured to operate like the outdoor
heat transfer unit 102 in FIGS. 1A and 1B described above as part
of a reversible refrigeration circuit also having a compressor,
condenser, reversing valve, and metering devices.
[0061] In FIG. 2, working fluid 113 enters a first heat exchanger
or first part 203 from an upstream condenser, such as indoor heat
transfer unit 111, at a first inlet 207. Working fluid 113 passes
through a first flow path 209 exiting at a first outlet 210
downstream from the first inlet 207 and upstream from a reversible
metering device 212 which operates like reversible metering devices
110 and 106 in FIGS. 1A and 1B. reversible metering device 212 has
an upstream inlet 211 and a downstream outlet 213, inlet 211
receiving working fluid 113 from first outlet 210, and outlet 213
passing the now lower pressure working fluid downstream to a second
inlet 214 defined by first part 203.
[0062] 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 FIG. 2, separate
first and second flow paths 215 and 209 place the warm condensed
working fluid 113 upstream from first part 203 and metering device
212 in a heat exchange relationship with the cooler, reduced
pressure working fluid 113 exiting metering device 212 and passing
downstream to second part 205. As illustrated, first flow path 209
and second flow path 215 flow in opposite directions in a
countercurrent fashion which can yield an increased rate of heat
transfer, however, countercurrent flow is not required. Similarly,
the illustrated embodiment has first flow path 209 passing inside
second flow path 215. In other embodiments, second flow path 215
may be inside first flow path 209, running alongside adjacent to
first flow path 209, or in less preferred embodiments, passing heat
across a gap between first flow path 209 second flow path 215.
[0063] 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 FIG. 2 operating in the
heating mode, second part 205 evaporates working fluid 113 by
exchanging heat between working fluid 113 moving through a second
part flow path defined by a conduit 219 arranged in a heat exchange
relationship with an external medium. For example, as illustrated
in FIG. 2, conduit 219 is arranged to traverse a tortuous path
through fins 221 increasing the surface area of for conduit 219 and
the corresponding rate of heat exchange between an external medium
outside conduit 219 and working fluid 113 inside. The external
medium may be any suitable heat source available such as a gas like
ambient air, or a liquid such as water or a solution or suspension
containing water and other chemicals, or other external media like
the ground.
[0064] After working fluid 113 evaporates to the vapor phase,
working fluid 113 exits second part 205 entering a downstream
compressor as illustrated in FIG. 1A. As discussed above with
respect to reversible metering devices 110 and 106, metering device
212 operates in a similar fashion and includes a sensor 223, such
as a temperature sensing bulb, for detecting the temperature of
working fluid 113 as it exits second part 205 in the vapor phase.
Temperature input to metering device 212 is provided through a
control line 225 which may be a wire, a conduit, a tube, or other
suitable connection. Metering device 212 responds to this input by
increasing or decreasing the flow of working fluid 113 entering
two-stage evaporator 200 accordingly.
[0065] Other embodiments of second part 105 from FIGS. 1A and 1B
are envisioned as well. For example, illustrated in FIG. 3 is
another embodiment of a multistage or multipart evaporator 300
having a first part or first heat exchanger 303, and the second
part or second heat exchanger 305. The first part 303 operates
similar to first part 203 in FIG. 2 and includes a first inlet 307
receiving working fluid 113 from an upstream condenser like indoor
heat transfer unit 111. As with first heat exchanger 203, first
heat exchanger 303 passes the working fluid 113 along a first flow
path 309 downstream to a first outlet 310 and into metering device
212. Metering device 212 is here coupled to a divider 313 which
separates the flow of working fluid 113 exiting metering device 212
so that a portion of it enters a second inlet 314 to pass along on
a second separate flow path 315. Thus first part 303 exchanges heat
between the cooler substantially liquid working fluid 113 passing
along second flow path 315 and the warmer substantially liquid
working fluid 113 passing along first flow path 309 like first part
203 in FIG. 2. Working fluid 113 exits downstream from metering
device 212 at a second outlet 316 to enter second part 305. Thus in
the embodiment illustrated at 300, working fluid 113 exchanges heat
between warm condensed working fluid 113 in a substantially liquid
phase upstream from metering device 212 and a portion of the
cooler, lower pressure, working fluid 113 downstream from metering
device 212. As a result, second part 305 also receives working
fluid directly from metering device 212 as well as from first part
303.
[0066] 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.
[0067] 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 FIG. 2, except that in second part 305,
multiple conduits 219 are arranged in conjunction with multiple
fins 221 to exchange heat with working fluid 113 along separate
flow paths defined by the conduits. Here each flow path receives
working fluid 113 from divider 313. Four independent flow paths
defined by conduits 219 are illustrated, however more or less paths
may be advantageous as well. Regardless of their number, working
fluid 113 exits conduits 219 into a single collector 320 before
exiting two-stage evaporator 300 to a downstream compressor such as
the compressor 107 shown in FIG. 1A.
[0068] 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.
[0069] A perspective view of one embodiment of two-stage evaporator
300 is illustrated in FIG. 4 where first stage 303 appears as a
tube-in-tube heat exchanger arranged vertically along one end of
the second part fin-and-tube heat exchanger 305. Working fluid 113
enters from a condenser at first inlet 307 and exits two-stage
evaporator 300 from collector 320 to enter a downstream compressor.
As with all figures illustrated herein, the arrangement,
positioning, location, and appearance of illustrated parts is not
meant to be limiting but only illustrative. For example, first
stage 303 in FIG. 4 appears as an elongate structure with a single
bend part way along its length. However, other arrangements are
envisioned as well, some of which will be illustrated further in
the following figures.
[0070] FIGS. 5A and 5B illustrate another embodiment of a
partitioned evaporator at 502 as part of a reversible heat pump
system 500 operating in the heating mode like the embodiments
illustrated in FIGS. 1 through 4 and discussed above. Heat pump
system 500 in FIGS. 5A and 5B includes embodiments of compressor
107, reversing valve 109, and fluid lines 112 carrying the working
fluid 113 from the compressor 107 downstream to the indoor heat
transfer unit 111 illustrated as a liquid to air heat exchanger
positioned within the enclosed space 114. Heat energy 116 is
radiated into enclosed space 114 from the hot compressed working
fluid 113 in the vapor phase into the enclosed space. The condensed
working fluid 113 then passes into first part 503 of partitioned
evaporator 502 downstream from heat exchanger 111 and upstream from
second part 505.
[0071] First part 503 has similar inlets and outlets creating a
similar heat exchange relationship between one or more flow paths
like those shown in FIGS. 3 and 4 with respect to first heat
exchanger 303. Working fluid 113 enters at first inlet 507 and
exits at first outlet 510 having exchanged heat with cooler fluid
passing out of metering device 212 and divider 313 and into second
inlet 514. Working fluid 113 continues out of second outlet 516 to
enter the second stage heat exchanger 505 through one or more
conduits 219. A second stage heat exchanger 505 illustrated as a
liquid-to-air exchanger like the one illustrated in FIG. 4. Working
fluid 113 then passes through one or more flow paths within the
second stage heat exchanger 505 exiting as a vapor through single
collector 320 to reenter compressor 107 thus completing the closed
refrigeration circuit in the heating mode. The heat pump system 500
also includes a fan 550 which in the heating mode is configured to
draw surrounding heat 117 from ambient air toward second stage heat
exchanger 505.
[0072] In the embodiments shown at 500 in FIGS. 5A and 5B, first
stage 503 and second stage 505 are physically separated but remain
coupled together by fluid lines 112. First part 503 may be any heat
exchange apparatus as discussed previously capable of changing heat
between the upstream working fluid coming from condenser 511 and
the cooler working fluid downstream from the metering device 212.
For example, first part 503 is shown as a tube-in-tube heat
exchanger wrapped or twisted into a helical shape. As previously
discussed, the illustrations in the figures are illustrative only
and therefore other heat exchange apparatus in varying shapes,
sizes, and configurations are envisioned as well for first part
503.
[0073] FIG. 5B shows a similar arrangement and configuration of
parts is shown in FIG. 5A except that FIG. 5B illustrates an
example of first part 503 positioned inside or outside partitioned
evaporator 502. No particular limitation is envisioned with respect
to the location of first part 503 as it may be positioned in any
suitable location. For example, it may also be advantageous to
position first part 503 inside the enclosed space 114 for ease of
maintenance or for other purposes.
[0074] It may also be apparent in FIGS. 5A and 5B is that first
part 503 may be coupled to an existing unpartitioned evaporator
thus modifying or retrofitting it to become a partitioned
evaporator. Using the embodiment illustrated in FIGS. 5A and 5B, an
unpartitioned evaporator may be converted to a partitioned
evaporator by uncoupling one or more fluid lines passing working
fluid 113 from metering device 110 to metering device 106, and by
also uncoupling at least one fluid line passing fluid downstream
from metering device 106 to one or more conduits 219. First part
503, which may be any suitable heat exchange device such as the
helical tube-in-tube exchanger illustrated, may then be interposed
between metering device 106 and 110 as illustrated and described
above. The outlet from reversible metering device 110 (near the
indoor unit 111 operating as a condenser) can then be coupled to
inlet 507, and outlet 514 can be coupled to inlet 211 of metering
device 212. To complete the closed refrigeration circuit, one or
more of the working fluid lines coupled to divider 313 may be
coupled to second inlet 514, and second outlet 516 may be coupled
to one or more of the conduits 219 leading into second part 505. In
this way, the closed refrigeration circuit may be reestablished
after converting an unpartitioned evaporator to a partitioned
evaporator 502. The conversion process may include other actions
such as removing the working fluid 113 from the reversible
refrigeration circuit prior to separating the connections in the
fluid lines 112 and then reintroducing the same or a different
working fluid back into the refrigeration circuit after the closed
circuit is reestablished.
[0075] 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 FIG. 6 is
illustrated another embodiment of the heat pump system operating in
the heating mode at 600 having a partitioned evaporator 503 to heat
an enclosed space 114. As with previous FIGS. 1 through 5B, FIG. 6
shows a reversible closed-circuit refrigeration circuit including a
compressor 107, reversing valve 109, and indoor heat transfer unit
111 operating as a condenser rejecting heat 116 into the enclosed
space 114. Also included are reversible metering devices 110 and
212 for regulating the pressure of working fluid 113 pending on
whether system 600 is operating in the heating or cooling mode as
discussed above.
[0076] 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 FIGS. 5A and 5B. Working fluid 113
enters first part 503 at a first inlet 507 exiting at a first
outlet 510 having undergone heat exchange with the working fluid
113 downstream from reversible metering device 212 entering at
second inlet 514 and exiting first part 503 at second outlet
516.
[0077] 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.
[0078] FIG. 6 illustrates an example of a secondary medium
providing the heat of evaporation for a heat pump operating in the
heating mode having a partitioned evaporator that uses an external
medium that is not air. In the illustrated embodiment, it may be
said that the external medium providing the heat of evaporation 117
is the earth 620 as illustrated. However, it can also fairly be
said that the external medium providing the heat of evaporation is
secondary fluid 611 circulating through the secondary loop, part of
which may be subsumed in yet another external medium such as the
earth, a pond or lake, and the like. It may also be said that
secondary fluid 611 and an external medium 620 together operate as
an external medium.
[0079] 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.
[0080] 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.
[0081] 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.
* * * * *