Abstrict The present invention provides an apparatus for dehumidifying air
supplied to an enclosed space by an air conditioning unit. The apparatus
includes a partition separating the interior of the housing into
a supply portion and a regeneration portion. The supply portion
has an inlet for receiving supply air from the air conditioning
unit and an outlet for supplying air to the enclosed space. A regeneration
fan creates the regeneration air stream. The apparatus includes
an active desiccant wheel positioned such that a portion of the
wheel extends into the supply portion and a portion of the wheel
extends into the regeneration portion, so that the wheel can rotate
through the supply air stream and the regeneration air stream to
dehumidify the supply air stream. A heater warms the regeneration
air stream as necessary to regenerate the desiccant wheel. The invention
also comprises a hybrid system that combines air conditioning and
dehumidifying components into a single integrated unit.
Claims I claim:
1. An apparatus for dehumidifying air supplied to an enclosed space
by an air conditioning system, the apparatus comprising: a) a housing
having an interior; b) a partition separating the interior of the
housing into a supply portion for containing a supply air stream
and a regeneration portion for containing a regeneration air stream,
wherein the supply portion has an inlet for receiving supply air
from the air conditioning system and an outlet for supplying air
to the enclosed space, and wherein the regeneration portion has
an inlet for receiving regeneration air and an outlet for discharging
regeneration air; c) a fan in air flow communication with the regeneration
portion for creating the regeneration air stream; d) a rotatable
desiccant wheel positioned such that a portion of the wheel extends
into the supply portion and a portion of the wheel extends into
the regeneration portion, so that the wheel can rotate through the
supply air stream and the regeneration air stream to dehumidify
the supply air stream; and e) a heat source capable of heating the
regeneration air stream as necessary to regenerate the desiccant
wheel as it rotates through the regeneration air stream.
2. The apparatus of claim 1 wherein the heat source is a direct-fired
gas burner.
3. The apparatus of claim 1 further comprising a mechanism for
modulating the heat source to regulate the temperature of the regeneration
air stream.
4. The apparatus of claim 1 further comprising a bypass damper
between the inlet and the outlet of the supply portion for controlling
the amount of supply air passing through the desiccant wheel by
selectively bypassing the desiccant wheel.
5. The apparatus of claim 3 further comprising a mechanism for
modulating the bypass damper to regulate the amount of supply air
passing through the desiccant wheel.
6. The apparatus of claim 1 wherein the desiccant wheel is sized
to handle a fraction of the air flow processed by the air conditioning
system.
7. The apparatus of claim 1 wherein the air conditioning system
comprises a compartment housing a condenser, the apparatus further
comprising a duct or opening connecting the regeneration inlet air
to the compartment that houses the condenser, whereby the regeneration
inlet air can be preheated by the condenser.
8. The apparatus of claim 1 further comprising a mechanism for
varying the rotational speed of the desiccant wheel to control the
amount of moisture removed from the supply air stream or heat transferred
to the supply air stream.
9. An apparatus for dehumidifying air supplied to an enclosed space
by a packaged heating, ventilating, and air conditioning (HVAC)
unit, the apparatus comprising: a) a housing having an interior;
b) a partition separating the interior of the housing into a supply
portion for containing a supply air stream and a regeneration portion
for containing a regeneration air stream, wherein the supply portion
has an inlet for receiving air leaving the HVAC unit and an outlet
for supplying air to the enclosed space, and wherein the regeneration
portion has an inlet for receiving regeneration air and an outlet
for discharging regeneration air; c) a rotatable desiccant wheel
having an axis of rotation positioned such that a portion of the
wheel extends into the supply portion and a portion of the wheel
extends into the regeneration portion, whereby the wheel can rotate
through the supply air stream and the regeneration air stream to
dehumidify the supply air stream; d) a mechanism for varying the
rotational speed of the desiccant wheel to control the amount of
moisture removed from the supply air stream and/or the amount of
heat transferred to the supply air stream; e) a bypass damper between
the inlet and the outlet of the supply portion for controlling the
amount of supply air passing through the desiccant wheel by selectively
bypassing the desiccant wheel; f) a fan for creating the regeneration
air stream; and g) a gas burner for heating the regeneration air
stream as necessary to regenerate the desiccant wheel as it rotates
through the regeneration air stream.
10. A hybrid air conditioning and dehumidifying apparatus capable
of controlling the temperature and humidity of air supplied to an
enclosed space, the apparatus comprising: a) a housing having an
interior; b) a partition separating the interior of the housing
into a supply portion for containing a supply air stream and a regeneration
portion for containing a regeneration air stream, wherein the supply
portion has an inlet for receiving air and an outlet for supplying
air to the enclosed space, and wherein the regeneration portion
has an inlet for receiving regeneration air and an outlet for discharging
regeneration air; c) a fan in air flow communication with the regeneration
portion for creating the regeneration air stream; d) a fan in air
flow communication with the supply portion for creating the supply
air stream; e) a cooling coil positioned in the supply air stream;
f) a rotatable desiccant wheel positioned downstream of the cooling
coil, such that a portion of the wheel extends into the supply portion
and a portion of the wheel extends into the regeneration portion,
so that the wheel can rotate through the supply air stream and the
regeneration air stream to exchange moisture and/or heat between
the air streams; and g) a heat source capable of heating the regeneration
air stream as necessary to regenerate the desiccant wheel as it
rotates through the regeneration air stream.
11. The apparatus of claim 10 wherein the heat source is a direct-fired
gas burner.
12. The apparatus of claim 10 further comprising a mechanism for
modulating the heat source to regulate the temperature of the regeneration
air stream.
13. The apparatus of claim 10 further comprising a bypass damper
between the inlet and the outlet of the supply portion for controlling
the amount of supply air passing through the desiccant wheel by
selectively bypassing the desiccant wheel.
14. The apparatus of claim 13 further comprising a mechanism for
modulating the position of the bypass damper to regulate the amount
of supply air passing through the desiccant wheel.
15. The apparatus of claim 10 wherein the desiccant wheel is sized
to handle a fraction of the air flow processed by the apparatus.
16. The apparatus of claim 10 wherein the apparatus further comprises
a compartment housing a condenser, the apparatus further comprising
a duct or opening connecting the regeneration inlet air to the compartment
that houses the condenser, whereby the regeneration inlet air can
be preheated by the condenser.
17. The apparatus of claim 10 further comprising a mechanism for
varying the rotational speed of the desiccant wheel to control the
amount of moisture removed from the supply air stream or heat transferred
to the supply air stream.
18. A hybrid packaged heating, ventilating, and air conditioning
(HVAC) and humidity control apparatus capable of controlling the
temperature and humidity of air supplied to an enclosed space, the
apparatus comprising: a) a housing having an interior; b) a partition
separating the interior of the housing into a supply portion for
containing a supply air stream and a regeneration portion for containing
a regeneration air stream, wherein the supply portion has an inlet
for receiving air and an outlet for supplying air to the enclosed
space, and wherein the regeneration portion has an inlet for receiving
regeneration air and an outlet for discharging regeneration air;
c) a regeneration fan in air flow communication with the regeneration
portion for creating the regeneration air stream; d) a supply fan
in air flow communication with the supply portion for creating the
supply air stream; e) a cooling coil positioned in the supply air
stream; f) a rotatable desiccant wheel having an axis of rotation
positioned such that a portion of the wheel extends into the supply
portion and a portion of the wheel extends into the regeneration
portion, whereby the wheel can rotate through the supply air stream
and the regeneration air stream to dehumidify and/or heat the supply
air stream; g) a mechanism for varying the rotational speed of the
desiccant wheel to control the amount of moisture removed from the
supply air stream or heat transferred to the supply air stream;
h) a bypass damper between the inlet and the outlet of the supply
portion for controlling the amount of supply air passing through
the desiccant wheel by selectively bypassing the desiccant wheel;
and i) a gas-fired heater capable of heating the regeneration air
stream as necessary to regenerate the desiccant wheel as it rotates
through the regeneration air stream.
19. A method of controlling the temperature and humidity of an
enclosed space, the method comprising the steps of: a) providing
an air conditioning system having a supply outlet; b) providing
an active desiccant module comprising: 1) a housing having an interior;
2) a partition separating the interior of the housing into a supply
portion for containing a supply air stream and a regeneration portion
for containing a regeneration air stream, wherein the supply portion
has an inlet for receiving supply air from the air conditioning
system and an outlet for supplying air to the enclosed space, and
wherein the regeneration portion has an inlet for receiving regeneration
air and an outlet for discharging regeneration air; 3) a fan in
air flow communication with the regeneration portion for creating
the regeneration air stream; 4) a rotatable desiccant wheel positioned
such that a portion of the wheel extends into the supply portion
and a portion of the wheel extends into the regeneration portion,
so that the wheel can rotate through the supply air stream and the
regeneration air stream to dehumidify and/or heat the supply air
stream; and 5) a heat source for heating the regeneration air stream
as necessary to regenerate the desiccant wheel as it rotates through
the regeneration air stream. c) connecting the supply inlet of the
active desiccant module to the supply outlet of the air conditioning
system; d) connecting the supply outlet of the active desiccant
module to the enclosed space; e) cooling and/or dehumidifying the
supply air stream by passing it through the air conditioning system;
f) dehumidifying and/or heating the supply air after it has passed
through the air conditioning system by passing it through the active
desiccant module while rotating the wheel through the supply air
stream and the regeneration air stream to exchange moisture and/or
heat between the air streams; and g) supplying the supply air leaving
the active desiccant module to the enclosed space.
20. The method of claim 19 wherein the active desiccant module
further comprises a bypass damper between the inlet and the outlet
of the supply portion, and wherein the step of dehumidifying the
supply air stream further comprises the step of controlling the
level of dehumidification by selectively bypassing the desiccant
wheel.
21. The method of claim 19 wherein the air conditioning system
comprises a compartment housing a condenser, the method further
comprising the step of preheating the regeneration inlet air by
drawing it from the compartment that houses the condenser.
22. The method of claim 19 further comprising the step of varying
the rotational speed of the desiccant wheel to control the amount
of moisture removed from the supply air stream and/or the amount
of heat transferred to the supply air stream.
23. A method of controlling the temperature and humidity of an
enclosed space, the method comprising the steps of: a) providing
a packaged heating, ventilating, and air conditioning (HVAC) unit
having a supply outlet; b) providing an active desiccant module
comprising: 1) a housing having an interior; 2) a partition separating
the interior of the housing into a supply portion for containing
a supply air stream and a regeneration portion for containing a
regeneration air stream, wherein the supply portion has an inlet
for receiving air leaving the HVAC unit and an outlet for supplying
air to the enclosed space, and wherein the regeneration portion
has an inlet for receiving regeneration air and an outlet for discharging
regeneration air; 3) a rotatable desiccant wheel having an axis
of rotation positioned such that a portion of the wheel extends
into the supply portion and a portion of the wheel extends into
the regeneration portion, whereby the wheel can rotate through the
supply air stream and the regeneration air stream to exchange moisture
and/or heat between the air streams; 4) a mechanism for varying
the rotational speed of the desiccant wheel; 5) a bypass damper
between the inlet and the outlet of the supply portion for controlling
the amount of supply air passing through the desiccant wheel; 6)
a fan for creating the regeneration air stream; and 7) a gas burner
for heating the regeneration air stream as necessary to regenerate
the desiccant wheel as it rotates through the regeneration air stream;
c) connecting the supply inlet of the active desiccant module to
the supply outlet of the HVAC unit; d) connecting the supply outlet
of the active desiccant module to the enclosed space; e) passing
the supply air stream through the HVAC unit; f) dehumidifying and/or
heating the supply air after it has passed through the HVAC unit
by rotating the wheel through the supply air stream and the regeneration
air stream to exchange moisture and/or heat between the air streams;
g) controlling the dehumidification of the supply air by selectively
bypassing the desiccant wheel; h) varying the rotational speed of
the desiccant wheel to control the amount of moisture removed from
the supply air stream and/or heat transferred to the supply air
stream, and i) supplying the air leaving the active desiccant module
to the enclosed space.
24. A method of controlling the temperature and humidity of an
enclosed space, the method comprising the steps of: a) providing
a hybrid air conditioning and dehumidifying apparatus comprising:
1) a housing having an interior; 2) a partition separating the interior
of the housing into a supply portion for containing a supply air
stream and a regeneration portion for containing a regeneration
air stream, wherein the supply portion has an inlet for receiving
air and an outlet for supplying air to the enclosed space, and wherein
the regeneration portion has an inlet for receiving regeneration
air and an outlet for discharging regeneration air; 3) a fan in
air flow communication with the regeneration portion for creating
the regeneration air stream; 4) a fan in air flow communication
with the supply portion for creating the supply air stream; 5) a
cooling coil positioned in the supply air stream; 6) a rotatable
desiccant wheel positioned downstream of the cooling coil, such
that a portion of the wheel extends into the supply portion and
a portion of the wheel extends into the regeneration portion, so
that the wheel can rotate through the supply air stream and the
regeneration air stream to exchange moisture and/or heat between
the air streams; and 7) a heat source capable of heating the regeneration
air stream as necessary to regenerate the desiccant wheel as it
rotates through the regeneration air stream. b) cooling and/or dehumidifying
the supply air stream by passing it through the cooling coil; c)
dehumidifying and/or heating the supply air after it has passed
through the cooling coil by rotating the desiccant wheel through
the supply air stream and the regeneration air stream to exchange
moisture and/or heat between the air streams; and d) supplying the
supply air leaving the active desiccant module to the enclosed space.
25. The method of claim 24 wherein the apparatus further comprises
a bypass damper between the inlet and the outlet of the supply portion,
and wherein the step of dehumidifying the supply air stream further
comprises the step of controlling the level of dehumidification
by selectively bypassing the desiccant wheel.
26. The method of claim 24 further comprising the step of varying
the rotational speed of the desiccant wheel to control the amount
of moisture removed from the supply air stream and/or the amount
heat transferred to the supply air stream.
27. The method of claim 24 wherein the hybrid air conditioning
and dehumidifying apparatus further comprises a compartment housing
a condenser, the method further comprising the step of preheating
the regeneration inlet air by drawing it from the compartment that
houses the condenser.
28. A method of controlling the temperature and humidity of an
enclosed space, the method comprising the steps of: a) providing
a hybrid heating, ventilating, and air conditioning (HVAC) and dehumidifying
apparatus comprising: 1) a housing having an interior; 2) a partition
separating the interior of the housing into a supply portion for
containing a supply air stream and a regeneration portion for containing
a regeneration air stream, wherein the supply portion has an inlet
for receiving air and an outlet for supplying air to the enclosed
space, and wherein the regeneration portion has an inlet for receiving
regeneration air and an outlet for discharging regeneration air;
3) a regeneration fan in air flow communication with the regeneration
portion for creating the regeneration air stream; 4) a supply fan
in air flow communication with the supply portion for creating the
supply air stream; 5) a cooling coil positioned in the supply air
stream; 6) a rotatable desiccant wheel having an axis of rotation
positioned substantially collinear or parallel to the partition
such that a portion of the wheel extends into the supply portion
and a portion of the wheel extends into the regeneration portion,
whereby the wheel can rotate through the supply air stream and the
regeneration air stream to dehumidify the supply air stream; 7)
a mechanism for rotating the desiccant wheel at a plurality of speeds;
8) a bypass damper between the inlet and the outlet of the supply
portion for controlling the amount of supply air passing through
the desiccant wheel by selectively bypassing the desiccant wheel;
and 9) a gas-fired heater capable of heating the regeneration air
stream as necessary to regenerate the desiccant wheel as it rotates
through the regeneration air stream. b) cooling and/or dehumidifying
the supply air stream by passing it through the cooling coil; c)
dehumidifying and/or heating the supply air after it has passed
through the cooling coil by passing it through the active desiccant
module while rotating the wheel through the supply air stream and
the regeneration air stream to exchange moisture and/or heat between
the air streams; and d) controlling the level of dehumidification
by selectively bypassing the desiccant wheel; e) controlling the
amount of moisture and/or heat transferred to the supply air by
adjusting the rotational speed of the desiccant wheel; and f) supplying
the supply air leaving the active desiccant module to the enclosed
space.
Description BACKGROUND OF THE INVENTION
[0002] The present invention pertains to the field of heating,
ventilating, and air conditioning ("HVAC"). More particularly,
this invention relates to systems and methods for controlling the
temperature and humidity of an enclosed space.
[0003] The quality of indoor air has been linked to many illnesses
and has been shown to have a direct impact on worker productivity.
New research strongly suggests that indoor humidity levels may have
a significant impact on the health of building occupants. For example,
microbes such as mold and fungus, which proliferate at higher indoor
humidity levels, have been shown to emit harmful organic compounds.
In addition to direct health effects, often the primary air quality
complaint of building occupants is unpleasant odors associated with
microbial activity. Building operators often attempt to eliminate
odors by increasing outdoor air quantities. This usually exacerbates
the problem because increasing outdoor air quantities often results
in higher indoor air humidity levels, which, in turn, fosters further
microbial activity.
[0004] The HVAC industry has responded to these indoor air quality
("IAQ") concerns through its trade organization, the American
Society of Heating, Refrigerating and Air-Conditioning Engineers
("ASHRAE"). ASHRAE Standard 62-1999 Ventilation for Acceptable
Indoor Air Quality, sets minimum ventilation rates and other requirements
for commercial and institutional buildings. Meeting these standards
generally requires systems capable of providing an increased supply
of outdoor air to the conditioned space while maintaining acceptable
humidity levels within the space. A large body of research supports
the need for continuous ventilation in accordance with ASHRAE 62-1999
while maintaining the relative space humidity between 30% and 60%.
IAQ problems including unacceptable odors and microbial infestation
often occur when HVAC systems fail to meet these design criteria.
[0005] Commercial and institutional facilities often use "packaged"
units, which combine air conditioning, heating and sometimes air
handling equipment in a single housing. Such systems are generally
designed to provide inexpensive heating and cooling. Such packaged
units are generally installed outside the building envelope, frequently
at ground level or on the building roof. A typical packaged unit
includes a supply fan and filter, a return air fan, a heating source
(typically an indirect gas fired heater or electric heating coil),
an outdoor air intake, and a mechanical refrigeration system consisting
of a compressor, cooling coil, and a condensing coil with a fan
that rejects heat to the outdoors. Typically a small fraction of
outdoor air is mixed with a much larger fraction of return air from
the building, conditioned by the unit then circulated through the
building by means of a system of supply and return ductwork. The
advantages of such packaged equipment include low purchase cost,
simplicity, familiarity, and compact design. More than 80% of all
air-conditioning systems sold to the commercial marketplace involve
compressorized package equipment.
[0006] A significant shortcoming of such packaged HVAC units is
that they are typically designed to utilize minimal outdoor air,
and, as such, are frequently incapable of handling the increased
continuous supply of outdoor air necessary to comply with ASHRAE
62-1999 guidelines. This is especially true in applications where
the need for 100% outdoor air systems exist, such as makeup air
to restaurants and hotel facilities. It is also true for applications
like schools, movie theatres and other facilities where a high occupancy
density results in the need for very high outdoor air percentages
being provided by the HVAC system.
[0007] To meet the increased outdoor air requirements of the ASHRAE
standards, HVAC professionals have attempted to use oversized packaged
equipment to match the increased cooling load associated with higher
outdoor air percentages. However, such oversized systems generally
suffer from sub-par performance and are expensive to operate. As
importantly, the oversized cooling capacity required to meet peak
outdoor air load conditions proves excessive at the more common
part-load conditions, and creates serious performance problems ranging
from over-cooling the space and lost humidity control due to reduced
compressor cycle times to freezing up coils and shortened compressor
life. Therefore, providing outdoor air continuously presents a tremendous
challenge to conventional packaged HVAC equipment.
[0008] For example, on mild, humid days (part-load conditions)
an oversized packaged unit will quickly cool the space to a set
temperature and then shut off the compressor. If the evaporator
fan is kept running to maintain a continuous flow of outdoor air
to the space, the indoor humidity level will usually climb due to
the humidity level of the outdoor air being introduced. This increase
in humidity will continue until the space temperature rises to the
point that the thermostat once again calls for cooling. By this
time, the humidity of the return air entering the cooling coil of
the packaged HVAC system is elevated. The elevated humidity of the
return air results in an elevated dew point temperature leaving
the cooling coil. Typically, the system can maintain space temperature,
but humidity control is lost, resulting in uncomfortable, cold,
clammy conditions. Occupants will often respond by lowering the
thermostat setting, causing the space relative humidity to further
increase. If such high humidity conditions persist, microbial growth
and other moisture-related IAQ problems may arise.
[0009] Another problem associated with oversized packaged equipment
selected to process outdoor air on a continuous basis results from
the re-evaporation of moisture that has condensed on the evaporator
coil. Henderson et al. (1998) and Khattar et al (1985) both have
confirmed the phenomenon, often observed in the field, where the
actual moisture removed by a packaged HVAC unit is significantly
less than anticipated based upon published performance data. Their
research shows that this reduction in dehumidification capacity
is attributable to moisture condensed on the direct expansion (DX)
coil evaporating back into the supply air stream when the coil is
cycled off but the fan continues to operate. Henderson (1998) has
shown that evaporation of moisture condensed on the DX coil can
reduce actual latent heat removal to less than 50% of the unit's
capacity at part load conditions. (1) Henderson, H. 1998. The Impact
of Part Load Air Conditioner Operation on Dehumidification Performance:
Validating a Latent Capacity Degradation Model. Proceedings ASHRAE
IAQ 98. (2) Khattar, M et. al. 1985. Fan Cycling Effects on Air
Conditioner Moisture Removal Performance in Warm, Humid Climates.
Presented at the International Symposium on Moisture and Humidity,
Proceedings. April, 1985 Washington D.C. (3) Henderson, H. 1990.
An Experimental Investigation of the Effects of Wet and Dry Coil
Conditions on Cyclic Performance in the SEER Procedure. Proceedings
of USNC/IIR Refrigeration Conference at Purdue University, West
Lafayette, Ind. July, 1990.) These and other limitations present
significant problems when packaged rooftop systems are forced to
handle high percentages of outdoor air volume, particularly if operated
as 100% outdoor systems. When applying a conventional packaged rooftop
system to handle all outside air, the cooling tons required at peak
conditions are far greater than the cooling output available at
the rated airflow of the conventional unit. This occurs because
standard conventional packaged cooling equipment currently available
on the marketplace by the major HVAC equipment manufacturers is
generally designed to accommodate only a relatively small portion
of outdoor air, typically 10-20%.
[0010] For example, a typical packaged gas/electric rooftop unit
available on the market today may have a rated cooling performance
at 95.degree. Fahrenheit (F) ambient, 80.degree. F. coil entering
dry bulb, 67.degree. F. coil entering wet bulb in accordance with
the ARI Standard 210/240-94. Assuming a typical ASHRAE/ARI outdoor
air cooling design condition of 95.degree. F. dry bulb and a 78.degree.
F. web bulb, and a return air condition of 78.degree. F. dry bulb
and 50% relative humidity, the mixed air condition entering the
cooling coil of 80.degree. F. and 67.degree. F. wet bulb corresponds
to an approximately 12% outdoor air percentage based on a simple
mixed air calculation.
[0011] Therefore, the design standard used to rate standard packaged
cooling equipment assumes that 80-90% of the air delivered to the
cooling coil is conditioned return air from the space. This return
air stream requires far less cooling capacity to condition than
raw outdoor air during peak cooling design conditions. As such,
the total cooling capacity needed by the standard conventional packaged
equipment would be greater if it were designed to accommodate a
much higher percentage of outdoor air.
[0012] For example, conditioning a 1500 cubic feet per minute
(cfm) outdoor air stream from 85.degree. F. and 130 grains (enthalpy
of 40.8 BTU/pound) to a 56.degree. F. dew point (enthalpy of 23.8
BTU/pound) requires approximately 10 tons of cooling capacity based
on a simple psychrometric calculation ((1500 cfm.times.4.5.times.(40.8-23.8)/12000
BTU/ton of cooling). However, the recommended minimum amount of
air capacity that can be processed by a typical 10 ton unit (alternative
1) without potentially causing problems such as frosting and compressor
failure is approximately 3000 cfm (300 cfm/ton). If the unit is
set up to provide 50% outdoor air (1500 cfm), and 50% return air
(1500 cfm) for a total of 3000 cfm across the cooling coil, the
cooling capacity must be increased to a 15 ton (alternative 2) unit
to accommodate the load associated with the extra 1500 cfm of recirculated
air. Problems such as coil frosting may be avoided in many cases,
since the mixed air temperature to the cooling coil is much closer
to the aforementioned design conditions of 80.degree. F. and 67.degree.
F. wet bulb. Examples of alternatives 1 and 2 are presented below.
[0013] If a standard 10 ton system is operated with only 1500 cfm
of air passing across the coil (only 150 cfm/ton), and if this air
is all outdoor air, the full 10 tons of cooling will be required
to reach a supply condition with a 56.degree. F. dew point. However,
when the outdoor air drops from the peak design condition of 95.degree.
F. and 78.degree. F. wet bulb to say 78.degree. F. and 64.degree.
F. wet bulb, the 10 ton compressor will deliver air as cool as 30.degree.
to 34.degree. F. At this point, the refrigeration pressure and temperature
will be very low, low enough to cause the moisture condensed on
the cooling coil to freeze. This frost buildup can result in increased
pressure loss across the cooling coil, which results in a reduction
of airflow, which results in more significant frost formation. This
and other problems associated with operating conventional DX cooling
systems at reduced airflow are well known to the industry and those
skilled in the art of refrigeration.
[0014] By applying a 15 ton system to process a total of 3000
cfm, 1500 cfm of which is outdoor air with the remainder being
return air, the mixed air condition to the coil is decreased from
the 95.degree. F. and 78.degree. F. wet bulb mentioned in the previous
example to approximately 86.5.degree. F. and 72.degree. F. web bulb.
At the peak condition, the 15 tons will provide a supply condition
having a dew point of approximately 56.degree. F. At the part load
condition used previously, 78.degree. F. and 64.degree. F. wet bulb,
the supply air condition will be approximately 40.degree. F. At
this condition, the refrigerant temperature is not as cold as the
previous example, and therefore may allow the coil to be operated
without freezing under part load condition.
[0015] However, using the increased cooling tons and supply airflow
may cause other operational problems. The higher 3000 cfm supply
airflow quantity may, for example, overcool the space, especially
at part-load conditions. This cooling causes the compressor to cycle
off, resulting in the delivery of high humidity air directly to
the space in addition to the moisture evaporated from the cooling
coil if the supply air fan continues to run. If, as a third alternative,
a 10 ton unit is used to process the 3000 cfm of total airflow,
of which 1500 cfm is outdoor air, at typical cooling season latent
design condition of 85.degree. F. and 130 grains, most conventional
packaged units of this size are only capable of delivering air at
a dew point of approximately 59.degree. F., even at a favorable,
return air condition of 75.degree. F. and 60% relative humidity,
and would therefore be incapable of maintaining the space at the
desired level of 50% relative humidity, since a dew point of approximately
55.degree. F. is required even if there is no latent load generated
by people or infiltration.
[0016] Customized overcooling reheat systems have been used in
an attempt to overcome these problems. However, such systems are
expensive to purchase and operate. Furthermore, complicated refrigeration
circuits frequently employed by such systems can be difficult to
troubleshoot and expensive to maintain. An example of the complexity
required to deliver a packaged piece of equipment to effectively
condition outdoor air is the TRANE.RTM. FAU product recently introduced
to the marketplace. The TRANE.RTM. Applications Considerations Bulletin
MUA-PRC004-EN shows a system that includes two separate evaporator
coils (an outdoor air evaporator and a main evaporator), three separate
condensing coils (a reheat condenser, a reheat outdoor condenser
and a main condenser), one reheat compressor, three main compressors,
two expansion valves, a subcooler and multiple complex controls.
[0017] Another attempt to meet the outdoor air and humidity level
requirements of the ASHRAE standards is through the use of "active"
desiccant-based systems, desiccant systems that employ a heated
regeneration air stream to remove moisture from the air. These active
desiccant systems have been used to reduce the humidity of outdoor
air prior to its introduction to the conventional HVAC system or
directly to the conditioned space. This allows the packaged equipment
to better control the space humidity despite increased outdoor air
requirements. Desiccants can be solid or liquid substances that
have the ability to attract and hold relatively large quantities
of water. In many commercial air conditioning applications where
desiccants are used, the desiccant is in a solid form and absorbs
moisture from the air to be conditioned. Examples of these types
of desiccants are silica gel, activated alumina, molecular sieves,
and deliquescent hygroscopic salts. In some cases, these desiccants
are contained in beds over which the air to be conditioned is passed.
Many times, however, the desiccant is contained in what is known
as an "active desiccant wheel."
[0018] An active desiccant wheel is an apparatus typically comprising
closely spaced, very thin sheets of paper, polymer film or metal
which are coated or impregnated with a desiccant material. The wheel
is usually contained in duct work or in an air handling system that
is divided into two sections: a supply section and a regeneration
section. The wheel is rotated slowly on its axis such that a given
zone of the wheel is sequentially exposed to the two sections. In
the supply section, the desiccant is contacted by the supply/outdoor
air. In this section, the desiccant wheel dehumidifies the supply/outdoor
air stream by absorbing moisture from the air onto its desiccant
surface. In the regeneration section, the desiccant contacts a regeneration
air stream (e.g., return/exhaust air being discharged from the space
or raw outdoor air). This regeneration air desorbs the moisture
from the desiccant that was adsorbed from the supply/outdoor air.
A heater is often used to heat the regeneration air stream as needed
to regenerate (i.e., dry) the desiccant wheel as it rotates through
the regeneration air stream. By cycling the wheel through these
two air streams, the adsorbing/desorbing operation of the wheel
is continuous and occurs simultaneously.
[0019] In the past, most active desiccant preconditioning systems
have not been coupled with rooftop packaged equipment, but applied
as stand alone systems. When they have been coupled with rooftop
packaged equipment, they have been positioned upstream of the packaged
unit in an attempt to control the humidity of the air entering the
conventional vapor compression system. Such systems have processed
the outdoor air by first passing it through an active desiccant
wheel to handle most of the latent load (humidity control), then
post-cooling the resulting warm, dehumidified outdoor air as necessary
to meet the temperature requirements of the conditioned space. However,
this approach generally has not found market acceptance because
of the relatively high purchase cost, high operational cost, large
size and inefficiency of such systems.
[0020] When an active desiccant dehumidification wheel removes
moisture from an air stream, heat is released as a result of the
adsorption process in addition to the heat contained within the
warm wheel media as it rotates from the hot regeneration air stream.
The more moisture absorbed, the more heat released. This heat significantly
increases the supply air temperature. In addition, removing large
quantities of moisture from outdoor air (e.g., 60 grains) requires
a high temperature air stream to regenerate the desiccant. In active
desiccant wheels, this high regeneration temperature is supplied
by an external heat source (e.g., a gas-fired heater). As mentioned,
the heat imparted to the desiccant wheel further increases the supply
air temperature. Based upon the literature for one of the best performing
commercially available active desiccant wheels, a 60 grain reduction
in outdoor air humidity would produce a 50.degree. F. increase in
the outdoor air temperature. Herein lies a significant problem with
the active desiccant preconditioning approach. If the desiccant
wheel handles all or most of the outdoor air latent load, the amount
of post cooling required to remove the sensible heat added by the
dehumidification process will often be similar to that required
to remove the humidity without the desiccant system. Consequently,
this approach generally does not reduce the overall system energy
consumption (total BTUs); rather, it increases it.
[0021] Another shortcoming of desiccant preconditioning approaches
attempted heretofore is that such systems have required very large
desiccant wheels to handle the significant latent load. For example,
to process only 1500 cfm of outdoor air, active desiccant wheels
as large as 42 inches have been applied. Including a standard cassette
and drive assembly, the height and width of the wheel unit required
is approximately 5 feet tall while a typical rooftop unit processing
the same amount of air is only 33 inches tall. Most prior active
desiccant systems have also employed a second, sensible only energy
recovery wheel to mitigate much of the process heat gained as a
result of the adsorption process. The size of the system required
to accommodate these two wheels, regeneration and other components
required is often four to five times the size of a comparable rooftop
package unit. These large systems are particularly undesirable for
commercial rooftop HVAC applications because they are more difficult
to install, require greater structural reinforcement, and are less
attractive. Architectural, engineering, economic and environmental
considerations all drive the desire to reduce the size and weight
of such packaged HVAC equipment.
[0022] Therefore, there is a significant need for energy-efficient,
compact HVAC system that can effectively control the temperature
and humidity of an indoor space while simultaneously providing high
quantities of outdoor air to the space. The present invention provides
these and other advantageous results.
SUMMARY OF THE INVENTION
[0023] The present invention provides systems and methods for controlling
the temperature and humidity of air supplied to an enclosed space.
[0024] An apparatus of the present invention for dehumidifying
air supplied by an air conditioning system includes a housing having
a partition separating the interior of the housing into a supply
portion and a regeneration portion. The supply portion has an inlet
for receiving supply air from the leaving side of the air conditioning
system cooling coil and an outlet for supplying air to the enclosed
space. The regeneration portion has an inlet for receiving regeneration
air and an outlet for discharging regeneration air. A fan in air
flow communication with the regeneration portion creates a regeneration
air stream.
[0025] The apparatus includes a rotatable desiccant wheel, which
is preferably sized to handle approximately 1/3 of the air flow
processed by the air conditioning system. The desiccant wheel preferably
positioned substantially collinear or parallel to the partition
such that a portion of the wheel extends into the supply portion
and a portion of the wheel extends into the regeneration portion.
The desiccant wheel rotates through the supply air stream and the
regeneration air stream to dehumidify the supply air stream. The
apparatus preferably includes a mechanism for varying the rotational
speed of the desiccant wheel to control the amount of moisture removed
from the supply air stream or heat transferred to the supply air
stream. The apparatus also preferably includes a bypass damper between
the inlet and the outlet side of the supply air portion around the
active desiccant wheel for controlling the amount of supply air
passing through the desiccant wheel. The bypass damper can also
be modulated to accommodate varying outdoor air and desired supply
air conditions by selectively bypassing the desiccant wheel.
[0026] A heat source (e.g., a direct-fired gas burner, indirect-fired
burner or heating coil) warms the regeneration air stream as necessary
to regenerate the desiccant wheel as it rotates through the regeneration
air stream. Heated air that is a byproduct of an air conditioning
system, a manufacturing process, and/or an electrical generation
plant, for example, may also serve as the regeneration source. The
apparatus can also include a duct or opening connecting the regeneration
inlet air to the compartment that houses the air conditioning condenser
to allow the regeneration heater inlet air to be preheated by the
condenser coil.
[0027] The present invention also includes a hybrid air conditioning
and dehumidifying apparatus and methods for using the apparatus
to control the temperature and humidity of air supplied to an enclosed
space. The hybrid unit includes a housing having a partition that
separates the housing into a supply portion and a regeneration portion.
The supply portion has an inlet for receiving air and an outlet
for supplying air to the enclosed space. The regeneration portion
has an inlet for receiving regeneration air and an outlet for discharging
regeneration air. A fan in air flow communication with the regeneration
portion creates the regeneration air stream and a fan in air flow
communication with the supply portion creates the supply air stream.
A cooling coil cools and/or dehumidifies the supply air stream.
A bypass damper can be positioned in the supply section to allow
a portion of the supply air leaving the cooling coil to bypass around
the active desiccant wheel, preferably allowing approximately 1/3
of the supply air flow to pass through the desiccant wheel under
normal operating conditions. A rotatable desiccant wheel positioned
downstream of the cooling coil further dehumidifies the supply air
stream. A portion of the desiccant wheel extends into the supply
portion and a portion of the wheel extends into the regeneration
portion, so that the wheel can rotate through the supply air stream
and the regeneration air stream to exchange moisture between the
air streams. A heat source heats the regeneration air stream as
necessary to regenerate the desiccant wheel as it rotates through
the regeneration air stream.
DRAWINGS
[0028] These, and other features, aspects and advantages of the
present invention will become more fully apparent from the following
detailed description, appended claims, and accompanying drawings
where:
[0029] FIG. 1 is a schematic top view of an apparatus for dehumidifying
air supplied to an enclosed space by an air conditioning system;
[0030] FIG. 2 is a partially broken away perspective view of an
apparatus for dehumidifying air supplied to an enclosed space by
an air conditioning system;
[0031] FIG. 3 is a schematic top view of a hybrid air conditioning
and dehumidifying apparatus capable of controlling the temperature
and humidity of an enclosed space;
[0032] FIG. 4 is a partially broken away perspective view of a
hybrid air conditioning and dehumidifying apparatus capable of controlling
the temperature and humidity of an enclosed space;
[0033] FIG. 5 is a diagram illustrating a sample of the expected
performance of a conventional desiccant-based air conditioning and
dehumidification system; and
[0034] FIG. 6 is a diagram illustrating a sample of the expected
performance of an air conditioning and dehumidification system in
accordance with the present invention.
[0035] For simplicity and clarity of illustration, the drawing
figures illustrate the general manner of construction, and descriptions
and details of well-known features and techniques are omitted to
avoid unnecessarily obscuring the invention.
DESCRIPTION OF THE PREFERRED EMBODIMENTS
[0036] The present invention applies a desiccant wheel in conjunction
with an air conditioning unit in a configuration designed to take
best advantage of the desiccant wheel and cooling coil to efficiently
produce very dry air, while minimizing the size and cost of the
combined system.
[0037] FIGS. 1 and 2 illustrate an embodiment of a system for controlling
the humidity of air leaving an air conditioning unit and entering
a conditioned space. An active desiccant module (ADM) 10 is positioned
to condition outdoor air downstream of the evaporator coil of a
standard HVAC unit 12. In the embodiment shown, the air conditioning
system is an HVAC unit. However, depending upon the system requirements,
an air conditioning unit (without heating and/or ventilating components)
can be used in place of the HVAC. The HVAC unit 12 can be any conventional
HVAC unit. In a preferred embodiment, the HVAC unit 12 is a standard
commercially-available packaged HVAC unit, for example, a TRANE.RTM.
VOYAGER.TM. rooftop unit. The HVAC unit 12 is preferably mounted
in a standard horizontal position. A short transition duct connects
the HVAC unit 12 to the inlet 14 of ADM 10. Supply air leaving the
HVAC unit 12 flows into the supply air inlet of the ADM 10 in the
direction indicated by the arrows on FIGS. 1 and 2. The supply fan
of the conventional HVAC unit 12 will, in most cases, provide the
desired airflow without the need for an additional booster fan.
However, the system can include a supplementary supply fan 17 if
desired.
[0038] The ADM 10 includes a housing 16 that surrounds the unit.
A partition 18 separates the unit into a supply air portion 20 and
a regeneration air portion 22. A regeneration air stream flows through
the regeneration portion 22 of the ADM 10 in the direction shown
by the arrows. A desiccant wheel 24 is positioned between the supply
air portion 20 and regeneration air portion 22. The wheel 24 preferably
has an axis of rotation 25 positioned substantially collinear or
parallel to the partition 18. The desiccant wheel 24 is positioned
to rotate through separate supply and regeneration air streams flowing
through the respective portions of the ADM 10. The wheel 24 preferably
includes a drive belt, which operates with a conventional drive
motor to rotate the wheel at a controlled speed. The wheel housing
preferably includes air seals to prevent air from escaping around
the edges of the rotating wheel 24. The ADM is preferably provided
with a main control panel housing the main system controls.
[0039] The desiccant wheel 24 may comprise any one of various devices
that removes latent energy (moisture) from one air stream and transfers
this latent energy to another air stream. In a preferred embodiment
the active desiccant wheel 24 is a rotary, desiccant coated, honeycomb
fluted matrix. The honeycomb matrix is made from a desiccant coated
substrate material such as very thin aluminum, fibrous paper or
polymeric materials to minimize conductivity and heat transfer.
In a preferred embodiment, the substrate material is evenly and
densely coated on both surfaces prior to being formed in the honeycomb
matrix to ensure that inner walls of the resulting flutes or channels
are essentially smooth thereby minimizing parasitic pressure loss
through the wheel matrix and maximizing the moisture storage capacity.
The preferred desiccant wheel utilizes a desiccant coating optimized
to provide the maximum amount of dehumidification or moisture adsorption/absorption
capacity when operated under moderate regeneration temperatures
ranging between approximately 175.degree. F. and 220.degree. F.,
although somewhat higher or lower regeneration temperatures may
also be used under some conditions. The desiccant coating should
also preferably adsorb or absorb moisture very effectively from
a cool, saturated air stream, then readily desorb the moisture when
the wheel media is rotated through the regeneration air stream.
The desiccant wheel preferably provides the desired moisture removal
at relatively high face velocities (greater than about 500 feet
per minute) through the active desiccant wheel matrix while minimizing
pressure loss (less than about 0.6 inches of water gauge).
[0040] The desiccant wheel 24 preferably is one having a very low-pressure
loss because it is advantageous to use a supply fan in the HVAC
unit 12 as the sole means of delivering air to the space. External
static capability is limited because most packaged units use forward
curve fans. The desiccant wheel 24 is preferably optimized for best
performance at moderate regeneration temperatures and with saturated
inlet conditions. A desiccant used for such a wheel desirably has
as high a water adsorption capacity as possible and therefore as
much useable desiccant mass on the wheel as is consistent with technical
and economic constraints (desirably, coating thickness of more than
one mil). Furthermore, although non-desiccant mass is required to
carry and support the desiccant material, the wheel preferably has
as little non-desiccant mass as possible because such mass increases
the weight of the wheel and reduces the wheel's dehumidification
capacity.
[0041] Desiccant materials may include, for example, A-type, X-type
or Y-type molecular sieves and other zeolites, various silica gels,
activated alumina, lithium chloride and other deliquescent salts,
hydrophobic polymers or other materials capable of adsorbing or
absorbing water vapor from an air stream. In a preferred embodiment,
a desiccant material that is capable of adsorbing/absorbing and
desorbing a high percentage of its own weight in water vapor while
processing a cool, humid air stream typical of that leaving a cooling
coil is desired. The preferred desiccant material should also operate
to provide the desired moisture removal capacity at design conditions
while utilizing a moderate regeneration air temperature ranging
from about 175.degree. F. degrees to about 220.degree. F., although
it is understood that desiccants requiring higher or lower regeneration
temperatures may also be configured to deliver acceptable conditions.
Finally, it is beneficial to minimize the amount of adsorption energy
generated as the moisture is absorbed/adsorbed onto/into the surfaces
of the desiccant material, so that the amount of heat and introduced
to the dehumidified air stream leaving the active desiccant wheel
can be kept to a minimum when desired.
[0042] Desiccant materials that have moisture isotherms that meet
these criteria include select Y type molecular sieves and most silica
gel desiccants and specifically larger pore, low density silica
gel powders that are capable of adsorbing a very high percentage
of their own weight when subjected to high relative humidity environments.
[0043] Laboratory and recent field test prototypes of the invention
have utilized an active desiccant wheel developed by SEMCO Incorporated,
which provides acceptable performance and meets the criteria outlined
previously for this component. The SEMCO model LT active desiccant
wheel employs a deep (270 mm depth) desiccant wheel media with relatively
large, sinusoidal flute openings having an approximate dimension
of 1.5 millimeters in height and 4.2 millimeters in width. This
media allows for the desired dehumidification performance while
meeting the low pressure loss criteria required to allow the existing
fan in the packaged rooftop unit to be capable of processing the
desired airflow through and around the active desiccant wheel without
the need for an additional booster fan in most cases.
[0044] The active desiccant wheel utilizes a very thin aluminum
base substrate material having a thickness of approximately 1.2
mils, coated on both sides with a composite mixture of a high surface
area Y type molecular sieve and silica gel desiccant materials.
This wheel is capable of providing the moisture adsorption capacity
and performance desired and as presented in Tables 1 and 2 below.
[0045] The desired performance has been obtained while utilizing
relatively high supply air face velocities through this particular
active desiccant wheel. For example, the standard 5 ton HVAC unit
tested to provide the performance data presented in Tables 1 and
2 utilized a wheel having a diameter of approximately 20 inches.
The net face area allocated to process the supply air stream was
approximately 0.93 square feet after compensating for the outer
rim, internal hub cover plate, the spokes and the seals. This wheel
processed approximately 36% of the 1400 total cfm, or 504 standard
cubic feet per minute (SCFM), during the testing completed to obtain
the performance data presented in Tables 1 and 2. Dividing the 504
standard cubic feet per minute processed by the 0.93 square feet
of net face area confirms a wheel face velocity of approximately
542 feet/minute.
[0046] The rotational speed of the active desiccant wheel 24 may
be adjusted to optimize the amount of dehumidification capacity
and/or reheat capacity sought. For example, in a preferred embodiment,
the speed of the desiccant wheel varies from a minimum of about
1/8 to about 1/2 rotations per minute (rpm) when in the dehumidification
mode and as high as about 8 rpm if used to provide winter heating
of the outdoor air. By modulating the rotational speed of the desiccant
wheel the amount of regeneration heat that is transferred to the
supply air stream can be varied.
[0047] At the lower speeds (e.g., 1/8 rpm), the carry-over heat
will be reduced. Reducing the amount of heat transfer is beneficial
when it is desirable to provide dehumidified air to the occupied
space that is as cool as possible. Such conditions typically arise
on warm, sunny days when the sensible space load is high.
[0048] At higher speeds (e.g., 1/2rpm), the carry-over heat will
be increased. Increasing the amount of heat transfer is beneficial
when it is desirable to provide dehumidified warm air to the occupied
space. Such conditions typically arise, for example, on cloudy,
rainy days when the space has a high latent load and a very low
sensible load.
[0049] Under most conditions, optimum dehumidification capacity
is achieved at an intermediate wheel speed of about 1/5 to 1/3 rpm.
At these speeds, the maximum amount of moisture is removed at most
of the conditions encountered during normal operation. The dehumidification
capacity will typically be reduced as the wheel speed is decreased
or increased appreciably from this intermediate speed. Particularly
when systems are built in accordance with this invention, and do
not choose to incorporate the modulating valve (variable regeneration
temperature capability) on the regeneration source, it will be advantageous
at times to reduce the speed of the active desiccant wheel below
this optimum range, sacrificing some of the dehumidification capacity
in exchange for the delivery of colder, less dry air from the system.
For example, during times when the space humidity is satisfied but
could benefit from additional cooling, the wheel speed would be
gradually reduced until either the space humidity was no longer
satisfied or the temperature within the space was as desired. Conversely,
it will be advantageous at times to increase the speed of the active
desiccant wheel above this optimum range, sacrificing some dehumidification
capacity in exchange for the delivery of warmer, less dry air from
the system. For example, during times when the space humidity is
satisfied but is cooler than desired, the wheel speed would be gradually
increased until either the space humidity was no longer satisfied
or the temperature within the space was as desired.
[0050] By adjusting the rotational speed of the desiccant wheel,
the system of the present invention can provide the further advantage
of providing supplemental heat for conventional HVAC units. During
the heating season, many standard packaged rooftop units do not
have enough heating capacity to accommodate high outdoor air percentages
on very cold days. Typical packaged units lack such heating capacity
because they are usually designed to process a minimal amount (about
15%) of outdoor air with most of the air entering the heater being
return air from the space. For example, a TRANE.RTM. 10 ton packaged
unit model YSC120A has a 202500 British thermal unit (BTU) output
indirect-fired gas heating section. At the rated airflow of 4000
cfm and with the outdoor air at 20.degree. F., the supply air temperature
will only be 67.degree. F., too cold to heat most spaces. By adding
an additional 73000 BTUs of heat using the active desiccant wheel,
the supply air temperature can be raised to 84.degree. F. This is
accomplished by raising the temperature of the 1350 cfm (1/3 of
the total airflow) passed through the active desiccant wheel by
50.degree. F. In order to avoid further dehumidification of the
outdoor air by the active desiccant wheel and to optimize the heating
efficiency function of the desiccant wheel (in this example, used
as an indirect-fired heat exchanger) the wheel speed can be increased
to the maximum setting of approximately 8 rpm.
[0051] The method and system of the present invention is particularly
useful when used in conjunction with conventional HVAC units having
electric heaters. A primary geographic market for the dehumidification
module described herein is areas where outdoor humidity levels are
typically high (often described as hot and humid climates). Such
markets typically do not experience long or extreme heating seasons.
Because heating requirements are minimal, packaged HVAC units sold
in these regions typically have electric heating coils. The electric
heating capacity provided is often not adequate for heating high
outdoor air percentages, even in mild climates. For example, the
standard TRANE.RTM. model TC061C3 has a maximum electric heating
output of 23 kW. This 5 ton unit, processing 2000 cfm, is therefore
capable of heating an outdoor air stream from 30.degree. F. to only
66.degree. F. The supplemental heating capacity offered by the dehumidification
module described herein facilitates use of electric heating despite
the high outdoor air percentages, and also allows for a possible
reduction in the size of the electric heating coil required.
[0052] In an embodiment of the invention, wheel speed reduction
and modulation is accomplished by a variable speed motor controller,
such as a frequency inverter, coupled with the motor driving the
active desiccant wheel. The drive motor can drive a belt around
the wheel, a friction wheel riding on the outer rim of the active
desiccant wheel, or be directly coupled to the shaft of the desiccant
wheel. A signal is provided to the frequency inverter from the system
control module or the building automation system to deliver the
desired supply air conditions.
[0053] The active desiccant wheel 24 is positioned downstream of
the cooling coil of the HVAC unit 12. As discussed in greater detail
below with reference to FIGS. 5 and 6 if the active desiccant wheel
were to be placed upstream (before) the cooling coil, the desiccant
wheel 24 would have to be much larger. A wheel positioned before
the cooling coil must process all of the outdoor airflow to reach
the desired moisture content leaving the active desiccant wheel.
Since typically more humidity would have to be cycled by the active
desiccant wheel located in this arrangement, and since the entering
outdoor air would almost always be at a lower relative humidity
than that leaving a wet cooling coil (near saturation), the velocity
of the air passed through the active desiccant wheel would likely
have to be lower than if the same wheel were installed after the
cooling coil as described herein.
[0054] When the active desiccant wheel is positioned to process
outdoor air before the cooling coil, it must remove far more pounds
of moisture than if it is positioned downstream of the cooling coil
as described herein, to produce the same desired supply air moisture
level to the conditioned space. In order to remove the much larger
moisture loads required, the active desiccant wheel positioned upstream
of the cooling coil must be operated at lower face velocities and/or
at much higher regeneration temperatures than is required by the
active desiccant wheel positioned downstream of the cooling coil.
The higher moisture loads and regeneration temperatures required
by active wheels installed upstream of the cooling coil results
in much more heat being added to the air leaving the active desiccant
wheel. Consequently, more energy is required to cool the air before
it is supplied to the conditioned space. Also, prior active desiccant
preconditioning approaches have required far more regeneration energy
than the system of the present invention because the desiccant wheel
processes more air and removes more moisture.
[0055] The desiccant wheel 24 is preferably sized to process approximately
33% of the air that passes across the cooling/heating coil of the
packaged HVAC unit 12. Sizing the active desiccant wheel to process
only a fraction of the total supply air stream is beneficial to
the overall size, performance and manufacturing cost of the active
desiccant module, three of the most important criteria for market
acceptance of this technology. Previous active desiccant systems
have been designed to process all of the outdoor air through the
active desiccant wheel. As a result, the size of the desiccant wheel
required by previous active desiccant systems is much larger than
required by the system described herein. If the size of the wheel
is larger, the overall size of the system is larger. Since the active
desiccant wheel has traditionally been the most costly component
in the overall system, the larger wheel results in higher manufacturing
cost.
[0056] The reduced size, manufacturing cost and increased energy
efficiency associated with the positioning of the active desiccant
wheel and the ability to process only a small fraction of the supply
air stream made possible by the present invention described herein
are only some of the more important and significant advantages offered.
Other equally important advantages, including, for example, improved
control options also exist.
[0057] Another significant advantage of this invention is that
the amount of bypass air can change from application to application
or within a given application to meet latent and sensible load requirements.
A modulating bypass damper 26 is positioned in the supply air portion
of the ADM 10 to maintain the desired flow through the desiccant
wheel 24. The supply air stream flows through the bypass damper
26 and/or desiccant wheel 24 to the conditioned space via outlet
15. Bypass damper 26 may be modulated from a completely closed position
to variable opened positions to control the flow through desiccant
wheel 24 or to completely bypass the desiccant wheel 24 during the
heating mode if desired. This configuration provides saturated air
to the desiccant wheel to maximize its operating effectiveness and
minimize the required regeneration temperature.
[0058] There are several advantages offered by this control option.
By moving more air through the desiccant wheel, dryer, warmer air
will be delivered by the system. By bypassing more air, cooler,
less dry air will be provided. The ability to modulate the bypass
air fraction allows the unit to cost effectively respond to changing
space sensible/latent load conditions, especially when the regeneration
energy is fixed. Another advantage offered by bypass air modulation
is that during a true "economizer" period, (when the outdoor
conditions are cool and dry enough to deliver directly to the space
without further conditioning) the desiccant wheel can be bypassed
to reduce the system internal static pressure and provide more outdoor
air to the space.
[0059] The preferred mechanism for modulating the damper is an
electric actuator. The percentage of open area of the damper, and
therefore the amount of air bypassing the active desiccant wheel,
is controlled by the installation of a modulating actuator such
as those manufactured by the Belimo or Siemans companies (for example
the Belimo model CM-24SR). A temperature and/or humidity sensor
can provide a signal to a control module. This control module can
be, for example, a direct digital control system or a simple combined
temperature sensor controller or space thermostat. The controller
processes the data from the sensors and sends a signal (for example
a 0-10 volts or 4-20 mA) to the actuator, causing it to open or
close the bypass damper to the extent necessary to provide the desired
supply air temperature and/or humidity conditions from the unit
or to maintain conditions within the conditioned space. Those skilled
in the art will appreciate that various other mechanisms for modulating
the damper are possible.
[0060] The desiccant wheel 24 is preferably an active desiccant
wheel. As used herein, the term "active desiccant wheel"
refers to a desiccant wheel that utilizes an external heat source
to regenerate the desiccant within the wheel media. In the examples
shown in FIGS. 1 and 2 the regeneration portion 22 of the ADM 10
includes a heat source comprising a heater 28 for regenerating the
desiccant wheel 24. Regeneration heat can be provide by any heat
source (e.g., gas, electric, hot water, steam, solar, waste heat
from air conditioning, mechanical or electrical generating systems,
etc.) capable of providing heat as required to regenerate (dry)
the desiccant wheel 24 as it passes through the regeneration air
stream. For example, heater 28 can be a direct-fired burner such
as an atmospheric line-burner. The heater 28 can also be a hot water
or steam coil, which may be preferred if waste heat is utilized
for regeneration of the desiccant wheel 24 if, for example, the
ADM 10 is used indoors or with a combined cooling, heating and power
(CCHP) system (where on-site power generation creates waste heat
as a by-product). Outdoor air is preferably drawn into the regeneration
portion 22 via regeneration air inlet 30 by a regeneration air fan
32 in air flow communication with the regeneration portion 22. As
used herein, the term "in air flow communication" refers
broadly to a fan or other air moving means positioned anywhere inside
or outside the apparatus so as to create the desired air steam.
The heated regeneration air stream flows through and regenerates
(dries) the portion of the desiccant wheel 24 rotating through the
regeneration portion 22. After passing through desiccant wheel 24
the regeneration air can be discharged to the outdoors via regeneration
air outlet 34.
[0061] The regeneration energy input can be modulated by a control
valve serving the direct fired gas burner to provide only the amount
of heat necessary to reach a desired dew point. A typical application
for this approach would be conditioning a school facility where
the desire is to provide a constant supply of dehumidified outdoor
air to each classroom conditioned to a specified dew point. As the
outdoor conditions change, the regeneration temperature is varied
until the desired delivered outdoor air condition is achieved. As
mentioned previously, the amount of regeneration energy can also
be modulated to avoid delivering air to that space that is cooler
than desired, even if the delivered dew point is being achieved.
Likewise, the energy input can be modulated when the active desiccant
wheel is used to provide a supplemental heating function.
[0062] There are many common ways to vary the energy input to a
heating device and those skilled in the art would be familiar with
these methods. The preferred regeneration source for the invention
is a direct-fired gas burner. The quantity of gas, and therefore
the heating output, is controlled by the installation of a modulating
butterfly valve in the gas piping prior to or "upstream"
of the burner module. This modulating butterfly valve is then opened
and closed as required by an actuator, of which many types are available
and known to those skilled in the art. A good example of which is
a rotary actuator manufactured by Eclipse Inc., model number ACT004.
A temperature and/or humidity sensor provides a signal to a control
module. This control module may be, for example, a direct digital
control system or a simple combined temperature sensor controller
or space thermostat. The controller processes the data from the
sensors and sends a signal (for example a 0-10 volts or 4-20 mA)
to the actuator, causing it to open or close the butterfly valve
to the extent necessary to provide the desired regeneration temperature
for the conditions encountered.
[0063] The regeneration energy input can also remain constant,
eliminating the added cost of the modulating valve and necessary
control components. The burner or other heat source can be cycled
much the same way a standard rooftop unit cycles both the cooling
coil and heating source. A dew point sensor can be placed in the
occupied space to control the cycling on and off of the regeneration
heater and the appropriate stages of cooling. When the dew point
sensor detects that space humidity is at a desired level, either
the DX section continues to run to provide further sensible cooling
if needed, based on the space thermostat setting, or if the space
dew point sensor and thermostat are satisfied, both the regeneration
heater and cooling coil stage or stages can be cycled off.
[0064] Hot gas or condenser heat can be used to augment the ADM's
regeneration energy requirement. When a packaged rooftop HVAC unit
is designed for optimum efficiency, the temperature of the condenser
heat is typically in the range of 125.degree. F. This heat creates
many "free" BTU's available for regeneration. This free
heat can be utilized, for example, by placing a second condenser
coil prior to the gas fired burner. However, this configuration
has the drawbacks of added equipment cost and controls complexity
associated with the addition of the second condenser coil, piping,
control valves and sensors. Also, the added pressure drop across
the second condenser coil located upstream of the gas burner increases
the cost of operating the regeneration fan and often increases the
size of the required fan motor. A preferred method of utilizing
this free heat is by drawing the regeneration inlet air from the
compartment that houses the condenser fan(s), condenser coil and
compressors to preheat the air entering the regeneration heater,
eliminating the need for a second condenser coil or controls modifications.
This configuration improves energy efficiency without increasing
equipment cost. This approach can increase the temperature of the
inlet air entering the regeneration heater by up to approximately
25.degree. F., providing a significant reduction in energy required
for regeneration heat.
[0065] The system of the present invention is particularly effective
when applied in conjunction with an HVAC system that is configured
to provide a high percentage of outdoor air to the conditioned space.
However, those skilled in the art will appreciate that the approach
can be effectively applied to applications requiring recirculated
air as well. Even 100% outdoor air systems often benefit from the
incorporation of a non-occupied mode. For example, if the system
of the present invention is used to condition outdoor air supplied
to a school, the classrooms will be unoccupied a very high percentage
of the time. All summer long, the school facility may not require
comfort cooling, but it nevertheless requires humidity control to
avoid microbial infestation. The system of the present invention
can be configured to allow for the reduction or elimination of outdoor
air delivery in lieu of recirculated air. By passing the recirculated
air through the ADM, the space humidity can be dehumidified to a
very low dew point, with minimum runtime. In addition, the dehumidified
air can be efficiently delivered at a room-neutral temperature to
avoid over-cooling spaces with minimal sensible load. The space
dew point can be monitored and the ADM activated only when the space
needs dehumidification.
[0066] In another embodiment of the present invention, the ADM
is configured as part of a fully integrated hybrid air conditioning
and dehumidifying unit. The hybrid unit preferably combines in a
single packaged system, the filtration, supply air fan, and DX evaporator
coil and condensing section and heating section and other components
typically found in the standard rooftop HVAC unit with the components
of the ADM as discussed above.
[0067] FIGS. 3 and 4 illustrate a hybrid air conditioning and dehumidifying
apparatus for controlling the temperature and humidity of an enclosed
space. The hybrid unit includes a housing 16 for containing the
apparatus. A partition 18 separates the housing into a supply portion
20 for containing a supply air stream and a regeneration portion
22 for containing a regeneration air stream.
[0068] The supply portion 20 has an inlet 14 for receiving outdoor
air. A supply fan 42 in air flow communication with the supply portion
20 creates the supply air stream. Supply air is drawn through supply
inlet 14 preferably through filters 44.
[0069] Cooling coil 46 positioned in the supply air stream cools
the supply air. Cooling coil 46 may be any of a variety of conventional
cooling devices, for example, direct expansion or chilled water
coils. In one embodiment, cooling coil 46 is a direct expansion
cooling coil, which is part of a conventional refrigeration system.
The compressor, condenser and condenser fan components of the refrigeration
system are housed in the condensing unit 48.
[0070] After passing through the cooling coil 46 the supply air
is dehumidified by desiccant wheel 24. Desiccant wheel 24 preferably
has an axis of rotation 25 positioned substantially collinear or
parallel to the partition 18 such that a portion of the wheel extends
into the supply portion 20 and a portion of the wheel extends into
the regeneration portion 22. The wheel 24 rotates through the supply
air stream and the regeneration air stream to dehumidify the supply
air stream. The system preferably includes a mechanism for varying
the rotational speed of the desiccant wheel 24 to control the amount
of moisture removed from the supply air stream or heat transferred
to the supply air stream. The system also preferably includes a
bypass damper 26 between the inlet 14 and the outlet 15 of the supply
portion 20 for controlling the amount of supply air passing through
the desiccant wheel 24 by selectively bypassing the desiccant wheel.
[0071] The regeneration portion 22 has an inlet 30 for receiving
regeneration air and an outlet 34 for discharging regeneration air.
A regeneration fan 32 is in air flow communication with the regeneration
portion 22 so as to create a regeneration air stream in the regeneration
portion. A heater 28 (e.g., a gas-fired burner and/or any other
heat source) heats the regeneration air stream as necessary to regenerate
the desiccant wheel 24 as it rotates through the regeneration air
stream. The regeneration air may be drawn from ambient air. Alternatively,
as discussed above in connection with FIG. 1 the regeneration air
can be drawn from the compartment 48 housing the condenser 50 to
provide regeneration air that has been preheated by the condenser
of the air conditioning system.
[0072] As those skilled in the art will appreciate, the various
components of the embodiments of the systems shown in FIGS. 1 2
3 and 4 can be placed in a variety of different configurations without
departing from the scope of the invention. For example, the hybrid
system shown in FIG. 3 has the condensing unit 48 on the left side
adjacent to the cooling coil 46 whereas in FIG. 4 the condensing
unit 48 is on the right side of the system separated from the cooling
coil 46. Various other modifications can be made to the illustrated
embodiments.
[0073] FIGS. 5 and 6 illustrate the advantage provided by the positioning
of the cooling coil before the desiccant wheel. FIG. 5 illustrates
the expected performance of a conventional desiccant preconditioning
approach in which the desiccant wheel is positioned before the cooling
coil. FIG. 6 illustrates an example of the performance and design
of a system having a desiccant wheel downstream of a cooling coil
in accordance with the present invention under the same conditions.
In both examples, the outdoor air conditions are 85.degree. F. and
125 grains of absolute humidity (68.5% relative humidity). It is
desired that each system supply 1400 standard cubic feet per minute
of air to the conditioned space at 77.5.degree. F. and 68.5 grains.
[0074] To achieve these supply air conditions, the conventional
system illustrated in FIG. 5 requires an active desiccant wheel
24 of the type described above having a diameter of approximately
34 inches (4.9 square feet of total net wheel face area after compensating
for the outer rim, internal hub cover plate, the spokes and the
seals). Assuming that regeneration air is supplied from the outdoors
under the temperature and humidity conditions described above, a
desiccant wheel of this size, processing this volume of air, would
require a 128500 BTU/hour heater 28 to produced approximately 800
SCFM of regeneration air at 225.degree. F. at an absolute humidity
of 90 grains. Under these conditions, the active desiccant wheel
24 would remove approximately 56.5 grains of humidity to achieve
the desired humidity level of 68.5 grains of moisture. Air entering
the cooling coil 46 would have a temperature of 136.degree. F. and
an absolute humidity of 68.5 grains. A cooling coil 46 of approximately
7.4 tons is required to reduce the temperature of the air to the
desired supply air condition of 77.5.degree. F. Under these temperature
and humidity conditions, the cooling coil 46 provides little or
no additional dehumidification.
[0075] FIG. 6 illustrates an example of the performance of a system
having a cooling coil 46 positioned to process air before the desiccant
wheel 24 as described herein. To achieve the desired supply conditions,
1400 SCFM of outdoor air is first passed through a conventional
ARI rated 5 ton packaged system providing 5.7 tons of cooling output
due to the high temperature and humidity conditions delivered to
the cooling coil 46 which cools the air to approximately 64.6.degree.
F. In contrast to the conventional approach described above, air
entering the cooling coil 46 is near saturation. Because the air
enters the cooling coil 46 at a higher relative humidity, the cooling
coil 46 is able to significantly dehumidify the air passing through
it, removing 38 grains of humidity. Because of the dehumidification
provided by the cooling coil 46 the desiccant wheel 24 need only
process 1/3 of the air to produce the desired supply air conditions.
Accordingly, the size of the wheel 24 can be significantly reduced.
In the example shown, the active desiccant wheel required by the
system is approximately 20 inches in diameter (compared with 34
inches required with the conventional approach). After subtracting
for the rim, hub and spokes, the net square face are of the 20 inch
wheel is approximately 1.55 square feet (compared with 4.9 square
feet for the wheel of the conventional approach). As a result the
size of the wheel installed prior to the cooling coil would have
to be 70% larger in diameter and 316% larger in area to perform
the same function. Assuming that regeneration air is supplied from
the outdoors under the temperature and humidity conditions described
above, a desiccant wheel of this size, processing the specified
volume of air would require a heater 28 having a capacity of 34992
BTU/hour (as compared with 128500 BTU/hour required by the conventional
system described above) to produced approximately 250 SCFM of regeneration
air at 200.degree. F. and an absolute humidity of 90 grains.
[0076] As moisture is cycled by an active desiccant wheel, heat
is released. The more heat released, the more the temperature rise
across the transfer media. In the conventional configuration presented
in FIG. 5 with the active desiccant wheel removing more moisture
(greater grain differential) from a much larger air stream, far
more heat is inherently added to the outdoor air stream prior to
the cooling coil. As a result, the cooling energy provided by the
cooling coil positioned after the active desiccant wheel is dedicated
to reducing the temperature leaving the active desiccant wheel to
the desired 77.5.degree. F. temperature. The cooling coil performs
essentially no dehumidification function (i.e., is operated as a
dry coil). In the example of a system in accordance with the invention
shown in FIG. 5 the cooling coil adds appreciably to the dehumidification
function and, since only approximately one third of the airflow
passes through the active wheel, far less heat is added to the supply
air stream. As importantly, much of the heat added is desirable
to bring the delivered air temperature to a room neutral condition.
[0077] Thus, it can be seen from the examples illustrated in FIGS.
5 and 6 that the system configured in accordance with the present
invention can produce the same desired supply air conditions with
a smaller desiccant wheel, smaller capacity cooling coil, and less
regeneration heat. As such, the present invention provides a system
that can occupy less space, consume less energy, and be manufactured
at a lower cost than conventional desiccant-based dehumidification
systems. These advantages are significant, particularly for packaged
rooftop HVAC applications where size and efficiency are paramount.
System Testing
[0078] An add-on ADM and an integrated hybrid system configured
in accordance with the invention were designed, produced, instrumented
and tested in an air test laboratory at the headquarters of SEMCO
Incorporated in Columbia, Mo. As those skilled in the art will appreciate,
the systems of the present invention can be configured in a variety
of ways. For testing purposes, the add-on ADM module was configured
with an active desiccant wheel produced by SEMCO, a direct fired
burner, regeneration fan, bypass damper, electrical package and
system enclosure as previously described. This module was coupled
via a short length (18" of insulated ductwork) with a standard,
single stage 5 ton TRANE.RTM. VOYAGER.TM. packaged rooftop unit.
The hybrid system was similarly configured except that the active
desiccant wheel and other components included in the ADM module
were integrated with all of the components included in the standard
5 ton VOYAGER.TM. packaged rooftop unit to form one homogenous system
resembling the standard rooftop in outward appearance (except it
was several feet longer) and not requiring the connecting ductwork.
The integrated hybrid system incorporated the regeneration energy
savings modification to pull the air entering the direct fired burner
from the condensing section comprising the TRANE.RTM. VOYAGER.TM.
compressor, condensing coil and condenser fan.
[0079] The test facility allowed the simulated outdoor air stream
entering the packaged HVAC unit and hybrid system to be carefully
conditioned and controlled to the desired temperature, humidity
and static pressure levels desired for performance monitoring. Both
the add on ADM module packaged unit combination and the hybrid system
were connected to the test facility and fully instrumented. All
instrumentation was connected to a central data acquisition system
and monitoring station, allowing real time data to be reviewed and
collected. The outdoor air conditions created by the facility's
preconditioning system were also controlled and maintained via a
direct digital control system (DDC), which was an integral part
of the test lab monitoring station.
[0080] A duct connection was made directly to the outdoor air intake
section of the HVAC unit (in the case of the ADM) and outdoor air
intake of the hybrid system. The condensing section ambient air
and the regeneration air were drawn from the test lab, which was
maintained at approximately 80.degree. F. and 90 grains absolute
humidity throughout most of the testing.
[0081] As expected, the performance was the same for both the ADM
operating in combination with a rooftop HVAC unit and the integrated
hybrid system. Tables 1 and 2 below summarize the key performance
parameters that were obtained from the testing.
1TABLE 1 Summary of test results showing system performance of
an ADM with a 5 ton rooftop HVAC unit and hybrid system. Outdoor
Estimated Actual Desiccant Air Cooling Coil Cooling Wheel ADM Test
Leaving Coil Leaving Leaving Leaving Regeneration Conditions Conditions
Conditions Conditions Conditions Temperature Dry Bulb Temperature/Grains
of Moisture (Deg. F.) 95.degree. F. 66.degree. F. 66.4.degree. F.
103.degree. F. 80.degree. F. 200.degree. F. 115 92 92 37 72 85.degree.
F. 65.5.degree. F. 64.6.degree. F. 100.5.degree. F. 77.5.degree.
F. 200.degree. F. 125 89 87 36 68 85.degree. F. 61.degree. F. 63.8.degree.
F. 99.5.degree. F. 77.degree. F. 200.degree. F. 110 76 84.5 31 65
95.degree. F. 63.5.degree. F. 64.2.degree. F. 101.degree. F. 77.degree.
F. 200.degree. F. 100 83 85 31 65 75.degree. F. 62.5.degree. F.
63.3.degree. F. 100.degree. F. 77.degree. F. 200.degree. F. 130
80 83.3 29 64 85.degree. F. 60.degree. F. 58.4.degree. F. 92.5.degree.
F. 71.degree. F. 200.degree. F. 90 73 69.7 23 53 75.degree. F. 56.degree.
F. 58.5.degree. F. 92.degree. F. 71.degree. F. 200.degree. F. 100
63 70 23 53 70.degree. F. 53.degree. F. 54.5.degree. F. 88.5.degree.
F. 67.degree. F. 200.degree. F. 90 57 60.4 20 46 65.degree. F. 53.degree.
F. 52.5.degree. F. 82.degree. F. 63.degree. F. 200.degree. F. 85
57 56.1 19 43 65.degree. F. Coil 65.degree. F. 101.degree. F. 78.degree.
F. 200.degree. F. 85 Off 85 29 65 90.degree. F. 57.degree. F. 57.9.degree.
F. 91.degree. F. 70.degree. F. 200.degree. F. 70 65 68.3 22 52
[0082]
2TABLE 2 Summary of test results showing latent performance of
ADM with a 5 ton rooftop HVAC unit and hybrid system. Latent Load
Processed Cooling Latent Load by ADM with Rooftop Outdoor Coil ADM
Processed by Unit or Integrated Air Leaving Leaving Rooftop Unit
Hybrid System Conditions Conditions Conditions Delivered Delivered
Dry Bulb Temperature/ Latent Dew Latent Dew Grains of Moisture Tons
Point Tons Point 95.degree. F. 66.4.degree. F. 80.degree. F. 1.8
65.6 F. 3.4 58.degree. F. 115 92 72 85.degree. F. 64.6.degree. F.
77.5.degree. F. 3.4 63.8.degree. F. 4.9 56.degree. F. 125 87 68
85.degree. F. 63.8.degree. F. 77.degree. F. 2.0 63.0.degree. F.
3.5 55.0.degree. F. 110 85 65 95.degree. F. 64.2.degree. F. 77.degree.
F. 1.2 63.4.degree. F. 2.7 55.0.degree. F. 100 85 65 75.degree.
F. 63.3.degree. F. 77.degree. F. 3.7 62.5.degree. F. 5.3 54.5.degree.
F. 130 83 64 85.degree. F. 58.4.degree. F. 71.degree. F. 1.6 57.6.degree.
F. 2.9 50.0.degree. F. 90 70 53 75.degree. F. 58.5.degree. F. 71.degree.
F. 2.4 57.7.degree. F. 3.7 50.0.degree. F. 100 70 53 70.degree.
F. 54.5.degree. F. 66.degree. F. 2.3 53.7.degree. F. 3.5 46.0.degree.
F. 90 60 46 65.degree. F. 52.5.degree. F. 63.degree. F. 2.3 51.7.degree.
F. 3.4 44.0.degree. F. 85 56 43 65.degree. F. 65.degree. F. 78.degree.
F. 0.0 64.2.degree. F. 1.6 55.0.degree. F. 85 85 65 90.degree. F.
57.9.degree. F. 70.degree. F. 0.1 57.1.degree. F. 1.5 49.0.degree.
F. 70 68 52
[0083] Parameters that were optimized in these studies include
the bypass air fraction, desiccant wheel speed and cfm/ton processed
by the cooling coil. By decreasing the bypass air fraction, drier
air could be delivered from the system, but this would result in
a slight increase in the delivered air temperature. By decreasing
the rotational speed of the desiccant wheel, a cooler delivered
air temperature could be obtained at the cost of some dehumidification
capacity. By reducing the cfm/ton of air processed by the cooling
coil, colder, drier air could be delivered at a reduced airflow
capacity. There are many parameters that can be adjusted for these
systems. Though such flexibility is an advantage as it relates to
control options of the systems of present invention, for testing
purposes, parameters must be fixed to allow performance data to
be displayed in a concise manner.
[0084] To simplify the presentation of the data, the systems were
operated as 100% outdoor air systems. For purposes of the test,
the standard 5 ton packaged HVAC unit was operated at 285 cfm/ton.
Approximately 64% of the supply air was bypassed around the active
desiccant wheel. Bypass fraction and regeneration temperature were
selected to achieve the delivery of preconditioned outdoor air at
a space neutral temperature (between about 68 and 78.degree. F.)
and at or below about a 57.degree. F. dew point (70 grains of moisture
per pound of dry air). The condenser temperature was maintained
at 80.degree. F. during the testing since it was located within
the laboratory facility. Regeneration inlet humidity conditions
were maintained at 90 grains. A constant regeneration temperature
of approximately 200.degree. F. was used to produce the data.
[0085] The 5 ton name plate rating of the tested system is associated
with the ARI testing criteria, but when a conventional packaged
rooftop unit is applied to deliver air to a cooling coil that is
warmer and more humid than that used for ARI ratings (for example,
85.degree. F. dry bulb and 76.degree. F. wet bulb as opposed to
the ARI 210/240-94 standard condition of 80.degree. F. dry bulb
and 67.degree. F. wet bulb), the actual BTUs of cooling capacity
delivered by the 5 ton unit is increased by approximately 10%, thereby
improving the overall system efficiency and highlighting another
advantage associated with the invention described herein.
[0086] Table 1 shows a wide range of outdoor air conditions. For
each of these outdoor air test conditions, the predicted values
for the coil leaving condition (based upon an interpolation of manufacturers'
data) in addition to the actual data measured during testing are
shown. Good agreement was found between the anticipated coil leaving
conditions and those recorded during laboratory testing. In addition
to the leaving coil conditions, the conditions leaving the desiccant
wheel and those supplied by the system are also presented.
[0087] The systems were tested twice at the 65.degree. F., 85 grain
outdoor air condition to highlight the ability of the systems to
handle outdoor air conditions that are cool yet humid, without the
need for operating the cooling coil compressorized section. This
highlights another significant advantage offered by the invention.
Coils within packaged DX cooling systems can reach a frosting condition
when processing high percentages of outdoor air during times when
outdoor air is cool and humid because the cool outdoor air conditions
allow the cooling cycle to produce far more tons than required.
The ability to cycle off the compressor during low-load conditions
reduces the risk of potential coil frosting and compressor failure,
thereby eliminating the need for costly control mechanisms. It also
significantly reduces energy consumption since the compressor can
be cycled off a large number of hours per year when the outdoor
air is cool yet humid.
[0088] The first 65.degree. F./85 grain test point shows the temperature
that would be delivered by the ADM/rooftop HVAC combination if both
were in operation. Air as dry as 43 grains (a 41.degree. dew point)
can be delivered at this condition. The second point shows how the
targeted supply air conditions are met without the use of any mechanical
cooling, only operating the ADM.
[0089] Table 2 provides the test data formulated in a different
way to highlight the increased latent capacity made possible by
the ADM. As shown, the ADM significantly increased the latent capacity
of the conventional 5 ton rooftop HVAC unit. The latent capacity
was increased by more than 88% at the latent design condition and
more than 125% at part load conditions without increasing the airflow
delivered or the amount of conventional cooling capacity utilized.
[0090] The ADM can be controlled and operated to deliver a variable
dew point with the regeneration input being constant. This control
method would be the most basic (least costly) control scheme. It
would also most closely resemble how packaged rooftop units are
typically controlled in the market today.
[0091] Just as the compressor is cycled on and off as the space
temperature or supply air temperature conditions are satisfied,
the ADM can be configured to simply cycle the regeneration burner
to deliver additional dehumidification until a present humidity
level is achieved. This control approach is the basis for the data
presented in Table 1.
[0092] The regeneration burner or other thermal regeneration source
can also be modulated and operated to deliver a desired dew point
to the space. If the system produces drier air than desired, the
amount of heat delivered to the regeneration air stream can be reduced
until the desired supply air condition is met.
[0093] Upon review of the test data shown in Tables 1 and 2 those
skilled in the art will appreciate that the systems configured in
accordance with the present invention allowed the conventional rooftop
unit to provide dry ventilation at room-neutral temperature in an
energy efficient manner. The very dry warm air leaving the desiccant
wheel mixes with the cooler, more humid bypass air leaving the evaporator
coil to produce the temperature and humidity condition desired.
System Comparison
[0094] There are numerous advantages offered by the present invention.
Some of those advantages are summarized below through a comparison
with the conventional packaged cooling approach and previously marketed
active desiccant systems.
[0095] Table 3 shows the results of a simple comparison made between
the ADM/rooftop packaged HVAC combination and a customized 8.5 ton
packaged unit designed to handle 100% outdoor air. The comparison
assumes that each system will process 1400 cfm of outdoor air from
a cooling season design conditions of 85.degree. F. and 125 grains
to a 56.degree. F. dew point. It also assumes that in order to avoid
over-cooling the space, the outdoor air will be reheated to 70.degree.
F. prior to its introduction to the space and assumes that 2 degrees
of fan heat exists. The energy analyses assume continuous operation
and use utility costs of $0.07/kWh for electricity and gas at $4.50/million
BTU.
3TABLE 3 Comparison of ADM/rooftop HVAC combination with standard
customized packaged rooftop unit. Rooftop HVAC with Custom DX Rooftop
ADM HVAC Unit Cooling Capacity 5 tons 8.5 tons Required Reheat Energy
Required 0 18145 (BTU/HR) Regeneration Energy 31050 N/A Required
(BTU/HR) Supply Dew Point Used 56.degree. F. 56.degree. F. for Analysis
Annual Cooling Energy $1360 $2315 Cost Approximate Unit Size 31"
.times. 46" .times. 46" 33.5" .times. 46.5"
.times. 83" (H .times. W .times. L)
[0096] As shown, the first obvious advantage is that the tons of
mechanical cooling required for the approach of the present invention
is only 59% that required by the customized packaged unit. Aside
from the obvious advantage of reduced electrical demand and electrical
service requirements, this reduces the amount of compressor cycling
since the smaller rooftop HVAC unit is fully loaded a much greater
percentage of the time. This also minimizes the problem of condensate
re-evaporation from the cooling coil discussed previously.
[0097] Table 3 also compares the estimated energy consumption for
both systems. As shown, this analysis projects the operating cost
of the ADM/rooftop HVAC combination to be 41% less than that of
the customized rooftop HVAC packaged unit (an HVAC unit having a
cooling capacity selected for processing all outdoor air, such as
manufactured by DECTRON.TM. or POOL-PACK.TM.). Designing a packaged
system to process 100% outdoor air involves the incorporation of
multiple compressors, hot gas bypass capacity controls, a cooling
coil that utilizes more rows than are typically applied to standard
rooftop units and a reheating coil which can be electric or hot
gas energy (hot gas used for the energy analysis shown in Table
3) resulting from the cooling cycle. The projected energy savings
could be greater in markets where gas rates are seasonally low during
the cooling season, where incentives are offered for gas cooling
and where electrical demand charges are high.
[0098] The ADM requires a secondary energy input for regeneration
that is not required by the customized rooftop HVAC. However, as
shown in Table 3 the regeneration energy at peak load conditions
is not significantly greater than the energy required for reheat
by the customized rooftop HVAC unit. If the regeneration inlet air
is pulled from the condensing section housing the compressors, the
resulting preheat can easily reduce the regeneration energy consumption
shown in Table 3 by more than 20%. More importantly, as the outdoor
air loads become less extreme, the amount of regeneration energy
required by the ADM can be reduced while maintaining the desired
supply air dew point. For the customized packaged unit, the amount
of reheat energy remains constant. The customized packaged unit
can be designed to use the heat of rejection from the refrigeration
circuit to provide "free" reheat. However, a tradeoff
in the reduction in the overall cooling efficiency (KW/ton) is required
to meet the reheat requirements. Additionally, the reheat temperature
delivered from a condensing coil of a refrigeration system is not
easily controlled. Another problem is that at part load conditions,
only one compressor for example may be required to dehumidify the
air to the desired dew point, so there may not be enough energy
generated by the cooling cycle to reheat to the desired supply air
temperature.
[0099] Another significant benefit of the system of the present
invention is that when the outdoor air is at cool and humid part
load conditions, the ADM allows the HVAC compressor to be cycled
off since all of the dehumidification needed can be provided by
the desiccant wheel. At these conditions, the customized packaged
unit requires the addition of hot gas bypass or multiple staging
with sophisticated controls to avoid frosting the cooling coil and
potentially damaging the compressor. During cool ambient conditions,
excess capacity is provided by the condensing section at the very
time that reduced capacity is required at the evaporator coil. Without
proper design considerations, this results in unacceptably low suction
temperatures (frozen coils). |