A method for conditioning air for an enclosure in which a supply
air stream is cooled with a refrigerant system containing a variable
compressor by passing the air over a cooling coil to reduce the
temperature thereof; the thus cooled supply air stream is then passed
through a segment of a rotating desiccant wheel under conditions
which increase its temperature and reduce its moisture content,
and then delivered to the enclosure. The desiccant wheel is regenerated
by heating a regeneration air stream with the condensing coil of
the refrigerant system, and then passing the heated regeneration
air stream through another segment of the rotating desiccant wheel.
At least one condition of the supply air stream, the regeneration
air stream, and/or the refrigerant system is sensed or monitored
and the output of the compressor is controlled in response to the
What is claimed is:
1. A method for condition air for supply to an enclosure comprising
the steps of cooling a supply air stream having a temperature range
of between 65.degree. F. 95.degree. and above and a moisture content
of between 90 180 gr/lb. with a refrigerant system cooling coil
to reduce the moisture content and temperature thereof to a first
predetermined moisture content saturation level and saturation temperature
range, passing the thus cooled and dried ambient supply air stream
through a segment of a rotating desiccant wheel under conditions
which increase its temperature to a second predetermined temperature
range of about 68 81.degree. F. and reduce its moisture content
further to a predetermined humidity level of between 30 80 gr/lb.;
and then delivering the thus treated air to said enclosure; regenerating
the desiccant wheel by heating a regeneration air stream with the
condensing coil of the refrigerant system to increase its temperature
to a predetermined temperature range of 105.degree. F. 135.degree.
F. and then passing the heated regeneration air stream through another
segment of the rotating desiccant wheel to regenerate the desiccant
in the wheel; sensing at least one condition of the supply air stream,
the regeneration air stream and/or the refrigeration system; and
controlling the output of the compressor in response to the sensed
2. The method as defined in claim 1 including the steps of supplying
make-up air to said supply air, sensing at least one condition of
the air in the enclosure and controlling the supply of make-up air
in response to such sensed condition.
3. The method as defined in claim 2 including the step of sensing
the regeneration air temperature entering the regeneration segment
of the desiccant wheel and controlling the volume of regeneration
air passing the condenser coil and entering the regeneration segment
of the condenser coil to control the air temperature entering that
segment to a predetermined value.
4. The method as defined in claim 3 including the step of sensing
the temperature of the cooled supply air leaving the desiccant wheel
and controlling compressor capacity in response to that sensed temperature
to maintain the cool air temperature leaving the wheel at a predetermined
5. The method as defined in claim 3 including the step of sensing
the condensing coil pressure and maintaining it at a predetermined
pressure condition, and controlling the volume of regeneration air
passing the condenser coil and entering the regeneration segment
of the condenser coil thereby to maintain a relatively uniform regeneration
6. The method as defined in claim 5 including the step of sensing
the temperature of the cooled supply air leaving the desiccant wheel
and controlling compressor capacity in response to that sensed temperature
to maintain the cool air temperature leaving the wheel at a predetermined
BACKGROUND OF THE INVENTION
Field of the Invention
The present invention relates to air conditioning and dehumidification
equipment, and more particularly to an air conditioning method and
apparatus using desiccant wheel technology.
It is well known that traditional air conditioning designs are
not well adapted to handle both the moisture load and the temperature
loads of a building space. Typically, the major source of moisture
load in a building space comes from the need to supply external
make-up air to the space since that air usually has a higher moisture
content than required in the building. In conventional air conditioning
systems, the cooling capacity of the air conditioning unit therefore
is sized to accommodate the latent (humidity) and sensible (temperature)
conditions at peak temperature design conditions. When adequate
cooling demand exists, appropriate dehumidification capacity is
achieved. However, the humidity load on an enclosed space does not
vary directly with the temperature load. That is, during morning
and night times, the absolute humidity outdoors is nearly the same
as during higher temperature midday periods. Thus, at those times
there often is no need for cooling in the space and therefore no
dehumidification takes place. Accordingly, preexisting air conditioning
systems are poorly designed for those conditions. Those conditions,
at times, lead to uncomfortable conditions within the building and
can result in the formation of mold or the generation of other microbes
within the building and its duct work, leading to what is known
as Sick Building Syndrome. To overcome these problems, ASHRAE Draft
Standard 62-1989 recommends the increased use of make-up air quantities
and recommends limits to the relative humidity in the duct work.
If that standard is properly followed, it actually leads to a need
for even increased dehumidification capacity independent of cooling
A number of solutions have been suggested to overcome this problem.
One solution, known as an "Energy Recovery Ventilator (ERV),"
utilizes a conventional desiccant coated enthalpy wheel to transfer
heat and moisture from the make-up air stream to an exhaust air
stream. These devices are effective in reducing moisture load, but
require the presence of an exhaust air stream nearly equal in volume
to the make-up air stream in order to function efficiently. ERVs
are also only capable of reducing the load since the delivered air
will always be at a higher absolute humidity in the summer months
than the return air. Without active dehumidification in the building,
the humidity in the space will rise as the moisture entering the
system exceeds the moisture leaving in the exhaust stream. However,
ERVs are relatively inexpensive to install and operate.
Other prior art systems use so-called cool/reheat devices in which
the outside air is first cooled to a temperature corresponding to
the desired building internal dew point. The air is then reheated
to the desired temperature, most often using a natural gas heater.
Occasionally, heat from a refrigerant condenser system is also used
to reheat the cooled and dehumidified air stream. Such cool/reheat
devices are relatively expensive and inefficient, because excess
cooling of the air must be done, followed by wasteful heating of
air in the summer months.
A third category of prior art device has also been suggested using
desiccant cooling systems in which supply air from the atmosphere
is first dehumidified using a desiccant wheel or the like and the
air is then cooled using a heat exchanger. The heat from this air
is typically transferred to a regeneration air stream and is used
to provide a portion of the desiccant regeneration power requirements.
The make-up air is delivered to the space directly, or alternatively
is cooled either by direct or indirect evaporative means or through
more traditional refrigerant-type air conditioning equipment. The
desiccant wheel is regenerated with a second air stream which originates
either from the enclosure being air conditioned or from the outside
air. Typically, this second air stream is used to collect heat from
the process air before its temperature is raised to high levels
of between 150.degree. F. to 350.degree. F. as required to achieve
the appropriate amount of dehumidification of the supply air stream.
Desiccant cooling systems of this type can be designed to provide
very close and independent control of humidity and temperature,
but they are typically more expensive to install than traditional
systems. Their advantage is that they rely on low cost sources of
heat for the regeneration of the desiccant material.
U.S. Pat. No. 3401530 to Meckler, U.S. Pat. No. 5551245 to
Carlton, and U.S. Pat. No. 5761923 to Maeda disclose other hybrid
devices wherein air is first cooled via a refrigerant system and
dried with a desiccant. However, in all of these disclosures high
regeneration temperatures are required to adequately regenerate
the desiccant. In order to achieve these high temperatures, dual
refrigerant circuits are needed to increase or pump up the regeneration
temperature to above 140.degree. F. In the case of the Meckler patent,
waste heat from an engine is used rather than condenser heat.
U.S. Pat. No. 4180985 to Northrup discloses a device wherein
refrigerant condensing heat is used to regenerate a desiccant wheel
or belt. In the Northrup system, the refrigerant circuit cools the
air after it has been dried.
The invention as described in our parent application Ser. No. 08/795818
is particularly suited to take outside air of humid conditions,
such as are typical in the South and Southeastern portions of the
United States and in Asian countries and render it to a space neutral
condition. This condition is defined as ASHRAE comfort zone conditions
and typically consists of conditions in the range of 73 78.degree.
F. and a moisture content of between 55 71 gr/lb. or about 50% relative
humidity. In particular, the system is capable of taking air of
between 85 95.degree. F. and 130 145 gr/lb. of moisture and reducing
it to the ASHRAE comfort zone conditions. However, that system also
works above and below these conditions, e.g., at temperatures of
65 85.degree. F. or 95.degree. F. and above and moisture contents
of 90 130 gr/lb. or 145 180 gr/lb.
As compared to conventional techniques the invention of the parent
application has significant advantages over alternative techniques
for producing air at indoor air comfort zone conditions from outside
air. The most significant advantage being low energy consumption.
That is, the energy required to treat the air with a desiccant assist
is 25 45% less than that used in previously disclosed cooling technologies.
That system uses a conventional refrigerant cooling system combined
with a rotatable desiccant wheel. The refrigerant cooling system
includes a conventional cooling coil, condensing coil and compressor.
Means are provided for drawing a supply air stream, preferably an
outdoor air stream over the cooling coil of the refrigerant system
to reduce its humidity and temperature to a first predetermined
temperature range. The thus cooled supply air stream is then passed
through a segment of the rotary desiccant wheel to reduce its moisture
content to a predetermined humidity level and increase its temperature
to a second predetermined temperature range. Both the temperature
and humidity ranges are within the comfort zone. This air is then
delivered to the enclosure. The system also includes means for regenerating
the desiccant wheel by passing a regeneration air stream, typically
also from an outside air supply, over the condensing coil of the
refrigerant system, thereby to increase its temperature to a third
predetermined temperature range. The thus heated regeneration air
is passed through another segment of the rotatable desiccant wheel
to regenerate the wheel.
It is an object of the present invention to treat outside supply
air at any ambient condition and render it to practically any drier
and cooler psychrometric condition with lower enthalpy.
Yet another object of the present invention is to provide a desiccant
based dehumidification and air conditioning system which is relatively
inexpensive to manufacture and to operate.
Another object of the present invention is to heat make-up air
while recovering enthalpy from a return air stream.
Yet another object of the present invention is to provide a desiccant
based air conditioning and dehumidifying system using single, multiple
and or variable compressors operating at the highest suction pressures
possible to produce stable operating conditions and enhanced energy
A further object of the present invention is to utilize the exhaust
air from the building as a regeneration air source. This air will
be at an absolute moisture condition substantially lower than ambient
air for a portion of the year. Using this air and adding heat from
the condenser coil will produce a better sink for process air moisture
In accordance with an aspect of the present invention the system
of the present invention includes an air conditioning or refrigeration
circuit containing a condensing coil, a cooling or evaporation coil
and a compressor and a desiccant wheel having a first segment receiving
supply air from the cooling coil of the refrigeration circuit to
selectively dry the supply air. A regeneration air path supplies
regeneration air to a second segment of the desiccant wheel as it
rotates through the regeneration air path. According to the invention
this system is modulated to provide a constant outlet air condition
from the process portion of the desiccant wheel over a wide range
of inlet conditions and volumes. Preferably the system uses variable
compressors whose output can be varied in response to air or refrigerant
conditions at predetermined points in the system. In one embodiment
the system may be operated in numerous different modes from fresh
air supply only to supply of simultaneous cooled and dehumidified
air. In addition a particularly simple and inexpensive housing structure
for the system of the invention is provided.
The above, and other objects, features and advantages of the present
invention will be apparent in the following detailed description
of illustrative embodiments thereof, which is to be read in connection
with the accompanying drawings, wherein:
FIGS. 1 1A and 1B are schematic diagrams of a first embodiment
of the basic system of the present invention;
FIG. 2 is a psychrometric chart describing the cycle achieved by
the embodiment of FIG. 1;
FIG. 3 is a psychrometric chart describing the cycle achieved by
the embodiment of FIG. 1 using a different control system.
FIG. 4 is a schematic view of another embodiment of the present
invention which is adapted to treat make-up air and recover enthalpy
from the return air stream.
FIG. 5 is a psychrometric chart showing the cycle achieved with
the system of FIG. 4 in the cooling only mode;
FIG. 6 is a psychrometric chart showing the cycle achieved with
the system of FIG. 4 in the dehumidification only mode;
FIG. 7 is a psychrometric chart showing the cycle achieved with
the system of FIG. 4 in the dehumidification and cooling mode;
FIG. 8 is a psychrometric chart showing the cycle achieved with
the system of FIG. 4 in an enthalpy exchange mode;
FIG. 9 is a psychrometric chart showing the cycle achieved with
the system of FIG. 4 in a fresh air exchange mode;
FIG. 10 is a schematic diagram of an embodiment similar to that
of FIG. 1 but utilizing two compressors;
FIG. 11 is an evaporator cross plot for the system of FIG. 10;
FIG. 12 is a schematic diagram similar to FIG. 1 showing yet another
embodiment of the invention using a reactivation temperature control
FIG. 13 is a schematic plan view of a housing structure for use
with the system of FIG. 1.
Referring now to the drawings in detail, and initially to FIG.
1 thereof, a simplified air conditioning and dehumidification system
10 according to the present invention is illustrated which utilizes
a refrigerant cooling system and a rotating desiccant wheel dehumidification
system. This system is a refinement of the system disclosed in our
parent application. In this case the system takes air at any ambient
condition and renders it to practically any drier and cooler psychrometric
condition with a lower enthalpy.
In system 10 the refrigerant cooling system includes a refrigerant
cooling circuit containing at least one cooling or evaporator coil
52 at least one condenser coil 58 and a compressor 28 for the
liquid/gas refrigerant which is carried in connecting refrigerant
lines 29. In use, supply air from the atmosphere is drawn by a blower
50 through duct work 51 or the like, over the cooling coil 52 of
the refrigerant system where its temperature is lowered and it is
slightly dehumidified. From there, the air passes through the process
sector 54 of a rotating desiccant wheel 55 where its temperature
is increased and it is further dehumidified. That air is then provided
to the enclosure or space 57.
Desiccant wheel 55 of the dehumidification system is of known construction
and receives regeneration air in a regeneration segment 60 from
ducts 61 and discharges the same through duct 62. The wheel 55 is
regenerated by utilizing outside air drawn by a blower 56 over the
condenser coil 58 of the air conditioning system. This outside air
stream is heated as it passes over the condenser coil and is then
supplied to regeneration segment 60 to regenerate the desiccant.
The regeneration air is drawn into the system and exhausted to the
atmosphere by the blower 56.
In this embodiment, compressor 28 is a variable capacity compressor
and preferably an infinitely adjustable screw type compressor with
a slide valve. As is understood in the art the volume through the
screws in such a compressor is varied by adjusting the slide valve
and thus the volume of gas entering the screw is varied. This varies
the compressor's output capacity. Alternatively a time proportioned
scroll compressor, a variable speed scroll or piston type compressor
may be used to circulate the refrigerant in line 29 through a closed
system including an expansion device 31 between the condenser coil
58 and the evaporator or cooling coil 52.
It has been found that by using a single non variable compressor
in refrigeration systems, the compressor does more work than needs
to be done with the results that the desired set point of the system
may be over shot. By using variable compressors as described the
system can modulate to provide a constant outlet condition over
a range of inlet air conditions and volumes. That is, the operation
of the compressor is controlled in response to one or more conditions.
As a result, for example, one can maintain a desired usable and
selectable humidity condition leaving the desiccant wheel by modulating
the compressor capacity.
Such modulation can be achieved by using more than one compressor
or variable compressors, such as the time proportional compressor
offered by Copeland, or variable frequency compressors which use
synchronous motors whose speed may be varied by varying the hertz
input to the motor, which causes variation in work output.
The refrigeration system described above can be modulated or controlled
to provide a constant outlet condition over a range of inlet conditions
and volumes. It allows the system to be used in make-up air applications
to meet requirements for ventilation, pressurization or air quality
(e.g., in restaurants where make-up air is required to replace kitchen
exhaust air). Thus control of the delivered make-up air volume can
be made dependent on pressure (through use of pressure sensors for
clean rooms and the like), CO.sub.2 content (through use CO.sub.2
sensors) to control quality, or based on occupancy (using room temperature
sensors). Such sensors would control make-up air volume using known
techniques to control, for example, the speed of blower 50 or air
diverter valves (not shown) in duct 51. The system, using the variable
compressor, can still be modulated to accommodate the variation
of temperature or humidity caused by the addition of make-up air
in order to maintain the desired environmental conditions.
According to this invention a desired delivered air temperature
and humidity level for the supply air to the enclosure or space
57 can be maintained within the ASHRAE comfort zone discussed above.
From those temperatures and humidity conditions the corresponding
wet bulb temperature can be determined, establishing the desired
conditions represented at Point 3 on the psychrometric chart of
FIG. 2. This wet bulb temperature is used as the target set point
for the cooling and drying of the supply air (whether it is return
air alone or mixed with make-up air as described above). Utilizing
the variable capacity of the compressor 28 the capacity of the
cooling coil 52 is controlled to maintain the supply air temperature
leaving the coiling coil at a temperature which will allow the conditioning
of Point 3 to be attained after the air passes through the process
segment 54 of the desiccant wheel. This temperature will be slightly
lower than the calculated wet bulb temperature of the desired delivered
air. Thus, as shown in FIG. 2 supply air (in this case ambient
air as shown in FIG. 1) which will typically have a temperature
range of between 65.degree. and 95.degree. F. DBT and above and
a moisture content of between 90 180 grains/lb. enters the cooling
coil 52 at 95.degree. F. Dry Bulb Temperature ("DBT"),
78.5.degree. F. Web Bulb Temperature ("WBT") and a moisture
content of 120 grains/lb. (Point 1 on FIG. 2). As the air passes
through coil 52 its conditions move along the dotted line in FIG.
2 from Point 1 at relatively constant humidity until it reaches
saturation and its humidity is then reduced with temperature along
the saturation line to Point 2 where it leaves the coil in a saturated
condition of between 50.degree. 68.degree. DBT and 30 88 grains/lb.
moisture content, in this case at 61.degree. DBT and 80.4 grains/lb.
The air then enters the process segment 54 of the desiccant wheel.
As it passes through the wheel the air is dried and heated adiabatically,
following the approximate path of the wet bulb line. It is further
dried to its leaving condition of between 68 81.degree. F. DBT,
50 65.degree. F. WBT, and 30 88 grains/lb. moisture content, in
this case at Point 3 of 77.degree. F. DBT, 61.5.degree. WBT and
57 grains/lb. Of course it is understood that the compressor is
operated in response to the temperature of the air leaving the cooling
coil at Point C in FIG. 1 to achieve the desired final air temperature.
The length of travel down the line from Point 2 to Point 3 depends
on the regeneration conditions of wheel 55. In accordance with this
invention the regeneration air temperature is increased to provide
a longer path down the wet bulb line, i.e., more drying, and reduced
to provide less movement, i.e., less drying. In this manner the
appropriate drying of the wheel also can be achieved so that the
supply air leaving condition (Point 3) will equal the intended design
As will be understood, given the capacity demanded from the cooling
side set point, the condensing coil 58 will need to eject varying
amounts of heat to the ambient air stream entering that coil depending
on conditions at Point E (FIG. 1). The variable heat flux entering
at Point E would, under normal conditions, result in an uncontrolled
regeneration temperature F entering the wheel 55. According to the
present invention the volume of air flow through coil 58 is varied
by the use of a bypass or exhaust fan 70 in order to achieve the
appropriate regeneration temperature entering wheel 55. This is
done by sensing the temperature of air entering the wheel and controlling
the fan 70 to selectively increase or decrease the volume of air
drawn through coil 58 with blower 56 in order to control the temperature
of air entering the wheel. Any unnecessary volume of air is then
dumped to the atmosphere by fan 70. Airflow is increased to reduce
the temperature and reduced to increase the temperature. The remaining
air is then drawn through the desiccant wheel to provide the appropriate
desiccant dryness required to achieve the desired drying results,
i.e., the movement from Point 2 to Point 3 in FIG. 7. By dumping
excess air passing coil 58 when the air quantity required to maintain
the desired regeneration temperature exceeds the air flow needed
to regenerate the desiccant total, energy is conserved by not exposing
the incremental air flow to the pressure drop associated with the
desiccant wheel. It also means a smaller blower 56 may be used.
This system allows compressor 28 to operate at the highest suction
pressure necessary to obtain the leaving air condition, i.e., the
temperature of air leaving the wheel 55. When this is done the compressor
operates against the minimum pressure ratio possible to produce
the intended result. Thus the performance of the cycle is maximized,
reducing energy consumption.
When it is required to obtain additional sensible cooling a secondary
cooling coil 52' may be used to further cool air leaving the desiccant
wheel. This coil may be supplied with refrigerant from the same
compressor 28. As shown in FIGS. 1A and 1B this additional coil
52' can be placed on either side of blower 50. In the position shown
in FIG. 1A, coil 52' allows for reduction in the supply air temperatures
after a slight rise in the air temperature occurring from its passage
through blower 50. In the position shown in FIG. 1B, coil 52' is
upstream of blower 50 in the case where the temperature increase
from the blower is immaterial. Since the cooling coil performs more
efficiently on the suction side of a fan this is the preferred embodiment
where added blower heat is not a factor.
As an alternative to the control system described above, control
also can be achieved without the calculation of wet bulb temperature
by controlling the capacity of the cooling side of the device to
provide the desired cooling capacity for the space, i.e., controlling
the compressor using the desired space temperature and allowing
the condensing side of the system to modulate accordingly. In this
case the volume of air drawn through the condenser 58 is controlled
to achieve the required regeneration temperature, within limits
of acceptable condensing pressure, and thus also achieve the required
regeneration capacity. The regeneration temperature is increased
to reduce outlet humidity ratio, and decreased to reduce drying
capacity, within acceptable pressure limits. This system is shown
in FIG. 3 wherein ambient air at Point 1 95.degree. F. DBT 78.5.degree.
F. WBT, 120 grains/lb. enters the cooling coil. It follows the dotted
line to the saturated curve as it passes the cooling coil to Point
2 at 50.degree. F. saturated and 64.60 grains/lb. This air then
enters the process segment 54 of the desiccant wheel. As the air
passes through the wheel it dries and is heated adiabatically following
the approximate path of the wet bulb line to Point 3 which is its
leaving condition at 69.degree. F. DBT; 52.degree. F. WBT, 30 grams/lb.
The combined effect of minimizing and controlling the precooled
temperature and regeneration temperatures as described above achieves
the target leaving conditions within the ASHRAE comfort zone.
The length of travel down the wet bulb line depends on the regeneration
condition. As noted above the regeneration temperature is increased
to provide a longer path down the line, or more drying, and is reduced
in order to produce less drying. In the alterative control system
first described the sensible cooling capacity is increased allowing
the equipment to provide cooling of the space.
FIG. 13 shows a schematic plan view of an air conditioning/dehumidifying
unit 10 according to FIG. 1 wherein the components bear the same
reference numerals. As seen therein the unit 10 is contained in
a housing 100 in an arrangement which eliminates the need for the
duct work 51 61 described above. Housing 10 is a rectangular box
like structure which defines an internal plenum 100 that is divided
by an internal wall 102 into plenum sections 104 106. The desiccant
wheel is rotatably mounted in wall 102 so that its process segment
or sector 54 is located in plenum 104 and its regeneration segment
60 is in plenum 106. Blower 70 is located at one side 108 of plenum
106 to draw supply air through apertures (not shown) in the opposite
side 110 over and through coil 58. That air flows over the compressor
28 to cool that as well and is discharged through apertures in wall
108 to the atmosphere.
Blower 50 is located in plenum 104 near the process segment of
wheel 55 in a sub plenum 112 defined by a wall 114 in plenum 104.
Blower 50 draws supply air through openings (not shown) in end wall
116 over and through evaporator coil 52 and then through the process
segment 54 into plenum 112. From there the supply air is discharged
through openings (not shown) in wall 110 at sub plenum 112 to the
enclosure of separate duct work leading to the enclosure 57.
Blower 56 is mounted in plenum 106 adjacent the downstream side
of the regeneration segment 54 of the desiccant wheel. A baffle
or other separating or channel means 118 is positioned in plenum
106 adjacent wheel 55 and extends part way towards wall 108. As
described above, blower 56 draws some of the air leaving coil 58
through the regeneration segment 60 of the desiccant wheel to regenerate
the wheel. The baffle 118 prevents recirculation of air leaving
the wheel from recirculating back around the wheel. That air then
either mixes with air being expelled from the plenum by fan 70 to
the atmosphere or it may be separately ducted, in whole or in part,
to the supply air line.
This structure has numerous advantages including its compact size,
elimination of duct work, and reduction in condenser and regeneration
fan/blower horsepower. It also eliminates the use for any anti-back
draft louvers on the condenser circuit.
Another embodiment of the invention is illustrated in FIG. 4. In
this embodiment the system is adapted to treat make-up air and recover
enthalpy from a return air stream. Return air is often available
in applications where fresh air is provided due to high space make-up
air requirements resulting from occupant capacity, and where a large
amount of air is not required for space pressurization for infiltration
load minimization. This type of design is typically used for schools,
theaters, arenas and other commercial spaces where humidity need
not be controlled to below normal level (such as is required in
supermarkets and ice rinks, which see energy and quality benefits
from lower humidity conditions.) Moreover such large spaces use
large volumes of air which have substantial heat value in them.
The system 80 of this embodiment comprises a cooling coil 52 for
treatment of an outdoor ambient supply air stream A followed by
a desiccant wheel 55 and blower 50 for conveying the supply air
stream to the space or enclosures. This air stream constitutes the
make-up air. The evaporator or cooling coil 52 is connected to a
plurality of DX refrigerant compressor circuits. This is illustrated
in FIG. 4 as two coils 52 52' and their associated compressors
28 and 28'. However it is to be understood that the cooling circuit
containing coil 52 and compressor 28 may consist of more than two
separately operable circuits containing separate coils and compressors.
A second or regeneration air stream E is drawn from the space 82
and is of a quantity approximately equal to 50 to 100% of the make-up
air in the first air stream A. This air first flows through the
condensing coil 58 then through the regeneration segment of desiccant
wheel 55 and is ejected from the enclosure to ambient. The refrigeration
circuit for this system is designed such that the required heat
rejected (i.e., given up) in the condenser to the air stream does
not exceed the heat carrying capacity of the second air stream between
its return air temperature and the maximum refrigeration circuit
condensing temperature of approximately 130.degree. F. The refrigerant
from this coil 58 is then used to cool the first (supply) air stream.
As also seen in FIG. 4 one or more additional compressors are connected
to the cooling coil of the supply air stream. These are sized to
provide the additional cooling capacity to take the ambient make-up
air stream from ambient conditions down to 57.degree. 63.degree.
F. These additional cooling circuits possess their own condensing
circuits that eject their heat directly to ambient. This is shown
in FIG. 4 at condenser 58' which treats ambient air drawn through
it by fan 70.
In this embodiment, desiccant wheel 55 is equipped with a drive
motor arrangement that enables the desiccant wheel to rotate selectively
at high revolutions, namely 10 30 rpm, and at low revolutions, namely
4 30 rph. In the high speed mode the desiccant rotor will act as
an enthalpy exchanger and will transfer latent and sensible heat
between the regeneration and make-up air stream. In the winter an
enthalpy wheel heats and humidifies the make-up air, and in the
summer it will cool and dehumidify.
The system of this embodiment can operate in five different modes.
As described hereinafter, the compressors and wheel speed states
are changed to adapt the performance of the system to the space
requirements. The system can run in any or a combination of the
five modes. The main five modes are: Cooling only mode; Dehumidification
only mode; Cooling and dehumidification mode; Enthalpy exchange
mode; and Fresh air mode.
Operation of this system in the cooling only mode is illustrated
on the psychrometric chart of FIG. 5. In this mode desiccant wheel
55 is not operated and only the number of compressors necessary
to provide sufficient cooling to the space are operating. However
the compressor 28' whose condenser coil 58 is in the return air
line is not operating since the wheel is not operating. Operating
in this manner, as seen in FIG. 5 ambient air in air stream A enters
the bank of cooling coils at the conditions of Point 1 at 95.degree.
F. DBT, 78.5.degree. F. WBT, and 120 grains/lb. moisture content.
As it passes through the cooling/evaporator coils it moves along
the dotted line to and then down the saturation curve to Point 2
at 65.degree. F. saturated, 92.8 grains/lb. The air has been cooled
and dehumidified at this point, but not necessarily to the ASHRAE
comfort zone since no dehumidification from the wheel occurs. Heat
absorbed in the condensing coil 58' is simply rejected to the ambient
air stream via the condenser and fan 70.
Operation of the system of FIG. 4 in the dehumidification only
mode is shown in the psychrometric chart of FIG. 6. In this mode
the desiccant motor is operated at low speed mode (i.e., 4 30 rph)
and the compressor 28' which serves the condensing coil 58 in the
return air stream E is operating to heat the regeneration air. The
other refrigeration circuits, including compressors 28 and coils
58', 52 are not operating. Thus, as seen in FIG. 6 ambient air
A enters the bank of evaporation coils at the conditions of Point
1 at 95.degree. F. DBT, 78.5.degree. F. WBT, and 120 grain/lb.
As this air passes coil 52 52' it is cooled in coil 52' along the
dotted line on the chart to and down the saturation line to Point
2 at 65.degree. F. saturated, 92.8 grains/lb. Because the desiccant
wheel is operating, air stream A is processed in the wheel where
it is dried and heated adiabatically following the approximate path
of the wet bulb line. It leaves the desiccant wheel and is supplied
to enclosure 82 at the conditions of Point 3 at 79.degree. F. DBT,
66.degree. F. WBT and 75 grains/lb.
In this example and in typical operation the regeneration air taken
from the space 82 by blower 56 will be at conditions of about 80.degree.
F. DBT an 67.degree. F. WBT, approximately the same condition as
the supply air stream of ambient air. This regeneration air (i.e.,
the exhaust air from the space) is passed through condenser coil
58 receives heat rejected from that coil and then flows through
wheel 55 to regenerate it. This is a substantial advantage, in this
condition of operation, over the use of ambient air alone to regenerate
the wheel since the exhaust air leaving the condenser coil will
have lower relative humidity than if ambient air was used. Thus
it will absorb more moisture from the wheel and improve desiccant
performance over what is achievable with outside air alone. After
passing the wheel it is vented to the atmosphere.
Operation of the system of FIG. 4 in the cooling and dehumidification
mode is illustrated on the psychrometric chart of FIG. 7. In this
mode, as in the dehumidification only mode, desiccant wheel 55 is
rotated slowly (4 30 rph) but additional cooling is provided by
the additional cooling circuit or circuits containing coils 58',
52 and compressor 28 which are operated, as they do in the cooling
only mode. In this case the cooling and dehumidification modes work
together. The first stage of refrigeration circuit containing coil
58 52' and compressor 28' also operate and provide the reactivation
Operating in this manner, supply air A (either all ambient or a
mixture of ambient and some return air) enters the bank of cooling
coils at Point 1 (FIG. 7) at 95.degree. F. DBT, 78.5.degree. F.
WBT, 120 grains/lb. It again follows the dotted line and down the
saturation line to Point 2 exiting coil 52'. Because the second
or additional stages of cooling circuits are operating the condition
of that air continues further down the saturation line arriving
at Point 3 after exiting the secondary cooling stage 52. At that
point the supply air stream conditions are 57.degree. F. saturated,
69.5 grains/lb.rh. This air then enters the process segment 54 of
the desiccant wheel 55 where it is dried and adiabatically heated.
It follows generally the path of the wet bulb line and leaves the
wheel at Point 4 at 74.degree. F. DBT, 58.degree. F. WBT, and 48
Operation of the system of FIG. 4 in the enthalpy exchange mode
is illustrated in the psychrometric chart of FIG. 8. This mode is
typically used in summer when the outside air is at a higher enthalpy
than the indoor air, or in winter when indoor enthalpy exceeds outdoor
In this case the desiccant wheel 55 is driven at high speed (10
30 rpm) and all the refrigeration circuits are off. As shown in
FIG. 8 in winter, when 100% outside air is used having the conditions
at Point 1 of 40.degree. F. DBT, 32.degree. F. WBT and 12.6 grains/lb.
passage of the air through the process section 54 of the wheel will
cause the conditions of the air exiting the wheel to move along
the dotted line from Point 1 to Point 2 at 52.5.degree. F. DBT,
44.5.degree. F. WBT, and 30.5 grains/lb. From that point a conventional
heater 80 can heat the air to the desired room temperature. The
exhaust air drawn from the heater is supplied to section 60 to transfer
heat and moisture thereto.
In the summer condition using 100% outside air at Point 5 82.5.degree.
F. DBT, 56.degree. F. WBT and 42 grains/lb. the system will operate
in a reverse manner by causing the air to move along the dotted
line from Point 5 to Point 6 i.e., to 80.degree. F. DBT, 61.5.degree.
F. WBT, 42 grains/lb., just at the ASHRAE comfort zone.
Using the system of FIG. 4 in its enthalpy exchange mode with 50%
ambient air and 50% return air will cause the air conditioning entering
the desiccant wheel process section 54 to move from Point 3 to Point
4 on FIG. 8.
The final, fresh air exchange mode of operation of the embodiment
of FIG. 4 is shown on the psychrometric chart of FIG. 9. In this
case all cooling circuits and the desiccant wheel are off, and only
the blowers are on to constantly replenish fresh air. As a result
the system delivers fresh ambient air without heat recovery, cooling
Preferably the compressors used in this embodiment are also of
the variable type to provide more efficient operations.
Yet another embodiment of the present invention is illustrated
in FIG. 10. The system of this embodiment is similar to that of
FIG. 1 except that two compressors 28 are used in the refrigeration
circuit. As seen in the evaporator cross plot of FIG. 11 for a representative
two compressor cooling circuit two operating conditions for the
system are possible depending upon whether one or both compressors
are operating. To minimize energy use, by increasing the coefficient
of performance (COP) of the system it is desirable to operate the
system at the highest suction pressures possible which permits the
desired space humidity and temperature conditions to be achieved.
Operating one compressor instead of two wherever possible also conserves
FIG. 8 shows two sloping lines rising to the right showing the
capacity in BTUH of one and two compressors versus saturated suction
temperature with the compressors operating at 100% capacity for
that temperature. The term saturated suction temperature means the
temperature of the coolant gas leaving the evaporator cooling coil
52 and entering the compressors.
The three lines which slope upwardly and to the left in FIG. 11
represent the suction temperature of the refrigerant gas when the
supply air stream is at one of three conditions noted on the graph
and shows the corresponding capacity of the compressors at each
temperature. Where the two sets of sloping lines cross, the evaporator
and compressor are operating at the same conditions and therefore
the most efficiency.
Typically multiple compressors (as well as variable compressors)
have been operated to cut in and out of operation based on either
fixed pressure points detected in the refrigerant line or based
on the temperature of the supply air leaving the evaporator/cooling
coil. In the present invention, using a humidity control unit (i.e.,
desiccant wheel), the space humidity error can be used to control
compressor operation. Thus "error" is the difference between
the actual humidity sensed in the room or space and the humidity
set point (i.e., the desired humidity level). This signal is then
used to reset the suction pressure cut in point for the second compressor.
If the error is large, which means humidity is not being reduced,
the reset action will move the suction cut in pressure to a lower
setting. On the other hand if the error is small, or the unit cycles
on or off rapidity, reset will increase the suction pressure cut
in. In this way the unit operates at the highest suction pressure
possible producing the most stable conditions and increased energy
A still further embodiment of the present invention is illustrated
in FIG. 12 which also allows operation of the unit in cooling or
dehumidification, or in both modes simultaneously.
Existing technology has traditionally controlled the discharge
pressure of refrigeration systems (i.e., the pressure of gas leaving
the evaporator or cooling coil) to prevent excessively low discharge
pressure during winter. One common technique of head pressure regulation
is to reduce condenser fan speed, which produces the beneficial
side effect of reducing the power needed to operate the fan.
For humidity control units reducing fan speed has the same effect
and benefit at low temperatures. However, because cooling applications
and the humidity control units as used in the present invention
have the ability to operate in cooling, dehumidification, or both
modes simultaneously, a variation on the industry-accepted practice
of pressure head regulation is needed.
When not limited by high outside ambient temperatures or a condenser's
particular design criteria it is desirable to maintain the discharge
pressure of the compressor at the equivalent of between 80.degree.
F. and 100.degree. F. saturated discharge temperature. The control
system of this embodiment will, in the cooling mode, optimize cooling
performance by setting the head pressure set point within this range.
Maximum efficiency is achieved at lower pressure ratios, which are
characterized by higher suction pressures and lower discharge pressures.
On the other hand a desiccant wheel humidity control unit relies
on creating a sufficient difference between the supply air's entering
relative humidity and the regeneration air's relative humidity.
This is the force driving moisture transfer in the desiccant wheel.
It also is beneficial to operate the refrigeration system across
the lowest pressure ratio possible. This means that higher suction
pressures and lower condensing pressures should be used. The system
of the present invention balances the performance of the entire
unit without giving preference to either the refrigeration system
or the desiccant system.
To accomplish this a humidity sensor 90 is placed in the regeneration
air stream, after the heating condenser coil 58. An exemplary target
RH value would be in the range of 10 to 30 percent RH. Assuming
that saturation of the cooled air leaving the cooling coil 52 is
achieved (Point 2 on the psychrometric charts) the space humidity
sensor in space 57 would reset the head pressure to attain a specific
RH sensed entering the wheel. The reset would be limited to keep
the head pressure within a predefined range of conditions. For example,
with R-22 refrigerant the range of head pressure limits would be
from 168 psig (90.degree. F.) to 360 psig (145.degree. F.). These
are generally accepted conditions of operation for known scroll
compressors. This achieves a range of leaving air temperatures from
the condenser coil or inlet to the wheel of 80.degree. F. to 140.degree.
F. and avoids drawing up condenser head pressures with attendant
loss of performance in the refrigeration system. Thus the compressor
would run at the lowest head pressure while still producing the
target relative humidity. The savings would be that the 45.degree.
F. leaving air temperature obtained with a head pressure of 260
psig reaches the target RH % at a lower pressure thereby reducing
compressor power input while increasing refrigeration capacity.
Another way of accomplishing the same result would be by utilizing
the differential or elasticity of reactivation outlet or differential
temperature to reactive inlet temperature. For example, the desiccant
wheel will presumably have a lower outlet air temperature when the
wheel is still wet. Conversely the outlet air temperature will begin
to climb when the wheel is fully reactivated, i.e., dry. The temperature
of the air on either side of the wheel could be detected by conventional
temperature sensors 92 and continuously monitored. When air increase
in reactivation inlet air temperature yields a nearly similar increase
in outlet air temperature it indicates that the energy is not being
used to displace moisture from the wheel and therefore that head
pressure should be reduced by appropriate control of the compression.
Alternatively the control could be set to maintain a target 20.degree.
F. differential in temperature across the wheel.
This system reduces lost energy by matching reactivation energy
to load to reduce reactivation temperatures which in turn reduces
head pressure that results in improved refrigeration performance.
Although illustrative embodiments of the present invention have
been described herein with reference to the accompanying drawings,
it is to be understood that the invention is not limited to those
precise embodiments, but that various changes and modifications
can be effected therein by those skilled in the art without departing
from the scope or spirit of this invention.