An air conditioning system with thermal storage that is suitable
for using solar energy. A solar collector used energy from the sun
to evaporate water from a desiccant fluid. The desiccant fluid is
then flows into a mass transfer device, which removes moisture from
an air stream. Calcium chloride is the preferred desiccant material
and can serve as an energy-storage medium. Electric or fuel backup
can be used with this system to regenerate the desiccant material.
In some embodiments an indirect evaporative cooler is added to provide
sensible cooling. A new desiccant cooling system that is specially
designed to work with the properties of this desiccant and meet
comfort requirements of conventional air conditioning.
1) A desiccant cooling system comprising: a. A desiccant cooler
capable of cooling a fluid to a temperature below the ambient wet-bulb
temperature, b. A quantity of desiccant liquid sufficient for providing
at least about two hours of operation of said desiccant cooler at
design capacity, c. Means for regenerating said desiccant liquid,
d. Means for storing said desiccant liquid.
2) A desiccant cooling system of claim 1 wherein said desiccant
cooler comprises: a. A desiccant-gas heat and mass exchanger which
acts to cool and dehumidify said gas, b. Means for cooling said
desiccant liquid, c. Means for circulating said desiccant liquid
through said desiccant-gas heat and mass exchanger, d. A water-gas
heat and mass exchanger which humidifies said gas, and e. Means
of circulating gas that has been previously dehumidified by said
desiccant-gas heat and mass exchanger to said water-gas heat and
mass exchanger so as to cool said water below the ambient wet-bulb
CROSS-REFERENCE TO RELATED APPLICATIONS
 This application is a continuation-in-part of application
Ser. No. 09/549329 entitled "Solar Air Conditioner" filed
on Apr. 14 2000.
FIELD OF THE INVENTION
 This invention is in the field of air conditioning, specifically
thermally driven air-conditioners with thermal that are capable
of accepting a thermal input from solar or off-peak electricity.
BACKGROUND OF THE INVENTION
 Solar air conditioning has great potential to reduce energy
use from air conditioning. Sunlight is most plentiful in the summer
when air conditioning is required.
 The problem is that existing solar technologies have not
produced systems that are economically competitive with conventional
electrically driven systems. Prior work with solar air conditioning
has not produced practical systems. Solar air conditioning systems
have used two basic approaches in an attempt to capture the sun's
energy for cooling--thermal and photovoltaic.
 The photovoltaic systems use photovoltaic panels to convert
solar radiation directly into DC electricity. Photovoltaic systems
have two major advantageous attributes: they can use conventional
electrically driven air-conditioning equipment which is widely available
and inexpensive with the addition of the solar panels that use an
inverter to produce AC power, and they can use the utility grid
for backup power during dark or cloudy periods.
 Unfortunately there are other attributes: the high cost
of manufacturing, the low conversion efficiencies, and the need
for a continual stream of photons to produce power, create three
major disadvantages. First electricity from solar cells is very
expensive because of the high cost of the solar panels. (Panels
for a residential air conditioner can cost well over $10000.) Second
the space needed for powering the air conditioning units is large.
And third the panels provide no energy storage, which creates a
need for use of grid based electricity at night and on cloudy day.
In fact, the peak output from the solar panels occurs around solar
noon, while peak air-conditioning load occurs several hours later,
resulting in a significant mismatch between supply of needed power
and demand. This mismatch greatly reduces the value of the system
in reducing peak power demand to the utility, demands which recently
deregulated markets is demonstrating are much more expensive to
meet than had heretofore been obvious. For off-grid locations, the
only viable energy storage system to match the provision of power
to times when demand is high (later in afternoon and at night) is
batteries. Batteries have a high first cost, require periodic replacement,
and normally use toxic and/or corrosive materials. These problems
have prevented the use of photovoltaic systems in other than a few
high-cost demonstration systems.
 Thermal systems use heat from the sun to drive an air conditioner.
Typical approaches use a high-temperature flat-plate collector to
supply heat to an absorption system. Systems with concentrating
collectors and steam turbines have also been proposed. Natural gas
or other fuel is used for backup heat. While thermal systems have
the advantage of eliminating the need for expensive photovoltaic
panels, they have attributes that produce major disadvantages.
 One problem is the high cost and large size of the solar
collectors. Flat-plate collectors running at about 190.degree. F.
(90.degree. C.) require double-glazing and selective surface to
achieve reasonable efficiency levels, which greatly increases the
collector cost. This high collector cost reduces the comparative
attractiveness of such systems to standard vapor compression systems
driven by grid electricity. Large collector size also reduces the
potential market size by eliminating many locations from possible
use of the systems.
 Furthermore, existing thermal technologies also suffer from
the poor COP of absorption systems, typically around 0.5. When combined
with a typical collector efficiency of 20 to 50%, this inefficiency,
besides creating a need for large collector areas, makes the whole
system much less economically and environmentally attractive.
 Another important problem introduced by the performance
attributes of current solar thermal air-conditioning concepts is
the high-cost and large size of high-temperature thermal storage.
Large thermal storage is required to reduce backup energy (typically
gas) that would be used much of the time when their was a mismatch
between demand for cooling and solar inputs. This mismatch is the
discrepancy between high solar input at noon and large demands for
cooling during late afternoon, at night, and on cloudy days. A related
problem with existing concepts for thermally driven solar cooling
is the need for significant power input for circulating pumps and
fans, which further reduces the possible energy savings.
 Together these attributes for current concepts for thermally
driven solar cooling imply that the large majority of their energy
input would come from the backup fuel and electrical input for fans
and pumps. In essence, these various problems mean that these solar
systems are effectively very expensive gas-driven systems.
 No commercially available or conceptually proposed system
has been demonstrated that has the attributes that would be needed
for commercially viable solar air conditioner. Commercial success
will require the system to have the following attributes: low first
cost (The market tends to be first cost driven so it is critical
that the cost and thus ultimate selling price not be too high.);
small collector area (critical to cost and to finding many locations
in which installation is practical); small storage size (The mismatch
between solar supply and cooling demand requires storage if the
system is not to become a glorified means for using fossil energy
and if it is to be practical to install in many locations--as well
as low cost to manufacture.); easy to incorporate backup capability
(Regardless of storage capacity, the ability of the system to meet
demand in extreme and unusual circumstances will be critical to
market acceptance; customers demand perfection and then some.).
 Evaporative coolers are a related technology with a long
history. Direct evaporative coolers are the simplest and most common.
They consist of a means for moving air over a wet pad. Water evaporates
from the pad and thereby cools and humidifies the air. They are
commonly used for comfort cooling in warm, dry climates such as
those found in the southwest U.S.
 Indirect evaporative coolers are more sophisticated. An
indirect evaporative cooler means that air is cooled by contact
with a dry surface that is in turn cooled evaporatively.
 Desiccant systems dry air for air-conditioning purposes.
A typical system uses a solid desiccant impregnated on a wheel of
corrugated metal or plastic.
 Some more obscure systems appear in the patent literature,
but each has its own problems. Patents RE 20469; 4660390; 4854129
describe regenerative indirect evaporative coolers that use a portion
of the air exiting the dry cooler as inlet air to the wet side.
RE 20469 describes a cumbersome arrangement of coils and cooling
towers this complicated and expensive. 4660390 describes another
system that uses tubes in a crossflow configuration to transfer
heat between a wet side and a dry side. 4854129 also uses a system
that uses a cooling coil with water from a cooling tower.
 Patent 5050391describes another option for the desiccant
system. This system uses solid desiccant material and a true counterflow
arrangement for the heat exchangers. It also has essentially a single
stage of cooling which limits it performance and its ability to
use inexpensive desiccant materials.
DESCRIPTION OF THE FIGURES
 FIG. 1 shows a basic embodiment of the invention.
 FIG. 2 shows a design of an indirect evaporative cooler
used in this invention.
 FIG. 3 is a schematic psychrometric chart that shows how
this cooler works.
 FIG. 4 is drawing of a simple air-lift pump that is preferred
for use in the invention.
 FIG. 5 shows a diagram of a combination evaporative-desiccant
cooler that is a component of this system.
 FIG. 6 shows another cooler configuration that uses a water
mist for evaporative cooling.
 FIGS. 7 8 and 9 shows solar collectors that may be used
in the invention.
 FIG. 10 shows another embodiment of the invention.
 FIGS. 11 and 12 are plots of temperatures through the coolers
used in the invention.
 FIG. 13 is an embodiment that can use exhaust heat from
a gas turbine for cooling inlet air to the turbine.
 FIGS. 14a, 14b, and 14c show details of heat exchanger design
that may be used in the invention.
 FIG. 15 shows another preferred embodiment that is suitable
for use as a dehumidifier.
 FIG. 16 is a preferred embodiment of the invention that
uses a counterflow liquid-to-liquid heat exchanger and counterflow
liquid-to-air heat exchangers.
 FIG. 17 is a preferred embodiment of the invention that
is suitable for producing cooled water.
 FIG. 18 is an alternate embodiment showing a single-stage
cooler that rotates.
 FIG. 19 is a multistage embodiment that is a assembled from
the stages as shown in FIG. 18.
SUMMARY OF THE INVENTION
 The present invention is a liquid desiccant cooling system
with thermal storage capability.
DESCRIPTION OF THE INVENTION
 Description a Preferred Embodiment: FIG. 1 shows a preferred
embodiment of the invention. A flow of desiccant fluid 1 is pumped
by pump 9 to a solar collector 15 that acts as a regenerator for
the desiccant fluid. The fluid trickles over a collector surface
2 in the form of a thin sheet 3. A cover 4 transmits solar radiation
11 which warms the desiccant fluid as it flows over the collector
surface. A flow of air 10 removes water vapor that evaporates from
the desiccant fluid. The concentrated desiccant 5 leaves the collector
and flows to a mass-transfer device 6 that allows the desiccant
to absorb moisture from an air stream 8. The mass-transfer device
is preferably a direct-contact exchanger similar to those used for
direct evaporative coolers and may also include a pump for recirculating
the desiccant liquid through the device to ensure good mass transfer.
A supply air fan 8 moves the moves the air stream through the mass-transfer
 An indirect evaporative cooler 14 cools the air stream
8 without adding moisture to it. A fan 12 draws a secondary air
stream 13 through the cooler. The secondary air stream may be exhaust
air from a building, ambient air, or a portion of the conditioned
air leaving the evaporative cooler or mass-transfer device. This
indirect evaporative cooler is optional and may be eliminated in
cases where no sensible cooling is required.
 Indirect Evaporative Cooler Design: FIG. 2 shows the one
heat exchanger system that is suitable for use as an indirect evaporative
cooler for this system. The cooler has two basic parts--mass transfer
means 50 and an air-to-air heat exchanger 51. The cooler is shown
without a top cover for clarity. Corrugated panels 20 for secondary
air are oriented so that corrugations run from side to side while
corrugated panels 21 for primary air have corrugations that run
from end to end. Channels 35 formed by the corrugations in panels
21 allow for free flow of air through the panels. Likewise similar
channels run in a perpendicular direction through panels 20. The
panels 20 are stacked alternately with panels 21 so that the channels
for each panel are perpendicular to the channels for the adjacent
panels. The outside surface of the sheets may be covered with an
adhesive or filler material to ensure good contact between the sheets.
For maximum durability the panels are preferably made of polypropylene
or polyethylene plastic. Metals, such as aluminum, are also possible
 Another option is to use corrugated cardboard and paper.
Waterproof adhesive material, such as one based on acrylic or linseed
oil, can coat the paper or cardboard and joins the layers together
to form a single unit. The advantage of cardboard or paper is its
very low cost. The disadvantage is that it is may be less durable.
One advantage of this system is that it is not possible to create
condensation within the heat exchanger, which allows the possible
use of cardboard in some applications. This is especially true in
desiccant applications since it is possible to keep both air streams
above their respective dew point temperatures even when outdoor
conditions are at 100% relative humidity (such as rain or fog conditions).
 The main fan 29 moves primary air stream 38 through a single
pass through channels 35. The stacked panels 20 and 21 form a heat
exchanger that cools the primary air without addition of humidity.
A portion of the cooled primary air splits off and forms a secondary
air stream 39 which is moved by secondary fan 30. The secondary
air first flows through multiple passes of the air-to-air heat exchanger
to cool the primary air stream. The direction of the secondary air
flow through the heat exchanger is shown by the dashed arrows.
 The passes of the secondary air stream are preferably arranged
in a counter crossflow configuration with a mass transfer means
ahead of each pass of the heat exchanger. The mass transfer means
is preferably a direct evaporative cooler. The direct evaporative
cooling sections 23 form U-bends that direct the secondary air through
each pass. As shown in the figure three triangular pieces fit together
to form two mitered elbows which make each U-bend. Pass dividers
27 would normally be included to prevent excessive leakage between
passes in the wet media in each pass. Housing 31 ensures that excessive
air does not leak in or out of the heat exchanger.
 The chief use of this system is as an evaporative air cooler,
but many other applications are possible. In addition to air, this
system can work equally well with any number of nonreactive gasses
such as nitrogen, carbon dioxide, inert gasses, etc. This system
can also be used as a heater. For example if a desiccant liquid
is substituted for water and the entering gas stream has a high
relative humidity, the system would act to heat the gas stream.
Volatile liquids other than water-based solutions can be used in
the system, but they are very expensive and may pose risks with
flammability or toxicity.
 The direct evaporative cooling sections need to be thoroughly
wetted to ensure good evaporation while minimizing mineral deposits.
In addition there is normally a large change in the wet-bulb temperature
from one end of the heat exchanger to the other, so that water circulation
between passes needs to be minimized to reduce undesirable heat
exchange. These factors make it desirable to use multiple water
circuits with multiple pumps.
 For large systems using multiple pumps does not introduce
a significant cost penalty, but for small systems multiple pumps
can add greatly to the cost. One possible solution is to have multiple
pumps that share a common shaft and motor. Seals separate the pumps
from each other to minimize leakage and heat transfer.
 FIG. 2 shows another possible option for circulating liquid
using air-lift pumps 42. Air pump 38 supplies pressurized air through
air line 36. Drain 37 removes water from the bottom of the direct
evaporative cooling sections 23. Air bubbles into the water to create
a pumping action. Extra water can be supplied to the pumps to make
up for that lost to evaporation or blow down.
 Other configurations of the air-to-air heat exchanger are
possible. For example instead of stacking corrugated panels on top
of each other, it may be possible on use spacers between the panels
that are oriented in the same direction. The spacer could separate
the passes of the secondary air and allow free flow of the secondary
air over the panels. In this configuration the primary air would
flow inside the channels of the panels. This alternative configuration
should reduce material cost and reduces thermal resistance of the
walls between the two air streams.
 Another configuration would simply stack sheets with spacers
to direct air flow. For example sheets of paper can be separated
by corrugated cardboard spacers. The spacers would be on the order
of 0.1 inches thick to form a flow channel for air. The orientation
of the spacers would alternate so that the air flow for the secondary
air is perpendicular to the that for the primary air. This arrangement
would use a minimum amount of material and is a simple design and
would be the preferred configuration for materials, such as paper,
that are easily glued together.
 Indirect Cooler Theory of Operation: FIG. 3 is a psychrometric
diagram showing how the idealized behavior of the system. For the
case of conventional direct evaporative cooler, the process start
at entering air 80 and follow the constant wet-bulb temperature
line 87 (which is also essentially a line of constant enthalpy)
and approach ideal exit condition 81 which is along saturation curve
88. For the new system used as a cooler there are two exit conditions,
the supply air 82 and exhaust air 83. The primary air stream follows
the line of constant absolute humidity 90 and approaches the saturation
condition at point 82. A portion of this air exits system as supply
air and the rest moves along the saturation line 88 as it is heated
and humidified until it approaches the ideal exhaust condition 83.
This exhaust condition is ideally at the intersection of the saturation
line 88 and the constant dry-bulb temperature line 91. The result
is a colder supply air temperature than is possible with a simple
direct evaporative cooler.
 For the case of a heater, the conventional direct contact
system would again follow constant wet-bulb line 87. The process
would start at the entering air condition 80 and approach ideal
equilibrium point 84 which is on the desiccant equilibrium curve
89. For the new system, there are again two ideal exit conditions,
the supply air condition 85 and the exhaust condition 86. As with
the cooler, the ideal supply air temperature is the point along
the constant absolute humidity line 90 that is in equilibrium with
the liquid. In both cases only a fraction of the primary air stream
needs to be exhausted, typically 30 to 50%, which leaves the rest
as supply air.
 Air-Lift Pump: FIG. 4 shows detailed drawing of an air-lift
pump that is suitable for pumping liquid desiccant and water. Air
pump 100 supplies pressurized air 103 to air line 101. The air pump
is preferably an aquarium pump or similar design. The air line 101
discharges inside water pipe 104 and creates a flow of air bubbles
102. The bubbles lower the average density of the fluid column which
causes the air and water mixture to move upward. This upward movement
draws intake water 105. On the top end a separator 108 allows outlet
air 107 and outlet water 106 to discharge from the pump in separate
flows. The advantages of this pump include low cost, simple design,
reliability, no moving parts. This pump is excellent for handling
small liquid flows with a small head requirement. A single air pump
can drive many air-lift pumps an thus create many liquid circuits.
 Evaporative-Desiccant Heat Exchanger: FIG. 5 shows a heat
exchanger that adds the use of a desiccant. This arrangement has
eight stages of cooling, 170 171 172 173 174 175 176 and
177. The incoming primary air 140 enters the rightmost stage 177.
It then flows through the eight stages in a straight path where
it is cooled and dehumidified. The exiting primary air 141 splits
into two flows. A secondary flow 142 goes back through the heat
exchanger in a counter-crossflow arrangement. The remaining primary
air 143 is supplied to the load.
 As shown in this figure each stage of cooling includes an
evaporative pad for cooling and humidifying the secondary air and
an air-to-air heat exchanger that transfers heat between the two
air streams. In addition some of the stages include desiccant, which
dries the primary air stream. The desiccant is preferably a liquid,
such as an aqueous solution of calcium chloride, lithium chloride,
lithium bromide, glycol, or similar material. Materials such as
sodium hydroxide and sulfuric acid have excellent physical properties
but are very corrosive and dangerous to handle. Calcium chloride
is very inexpensive, has acceptable thermodynamic properties, has
relatively low toxicity, and is generally the preferred desiccant
material for this system.
 Starting with the rightmost stage 177 secondary air flows
over evaporative pad 158 and through air-to-air heat exchanger 159.
The primary air flows through the other side of the air-to-air heat
exchanger 159. The flow directions of the two air streams are perpendicular
to each other with the primary air going is a straight line through
the heat exchanger.
 Next to the left is stage 176 which includes a desiccant
161 that is in the primary air stream. The desiccant is preferably
provides a surface that is wetted by a liquid desiccant that is
in direct contact with the primary air stream. The evaporative pad
56 cools and humidifies the secondary air stream. The secondary
air stream cools the primary air stream in air-to-air heat exchanger
 The stages 170 175 173 and 171 are similar to stage 177
with evaporative pads 154 150 146 and 144 in the secondary air
stream and air-to-air heat exchangers 155 151 147 and 145 transferring
heat between the two air streams.
 The stages 174 and 172 are similar to stage 176. They include
desiccants 162 and 163 which dehumidify and increase the temperature
of the primary air stream. The evaporative pads 148 and 152 cool
and humidify the secondary air which cools the primary air in air-to-air
heat exchangers 149 and 153.
 Mist Cooling Option: FIG. 6 is another heat exchanger configuration
that uses a mist cooling system. Fan 220 draws in the main air stream
200. The air moves in a straight line through interior channels
209 in panels 208 and is cooled by evaporating water mist 205 on
the outside of panels 208. The mist is supplied by nozzles 204 that
are connected by way of pipe 212 to a source of pressurized water
213. The water is preferably demineralized and filtered to prevent
clogging and fouling of the nozzles and the heat exchanger surfaces.
As the main air stream leaves the heat exchanger, a portion of the
air forms a secondary air stream 202 which returns on the wet side
of the heat exchanger. Dividers 207 direct the secondary air in
multipass counter crossflow arrangement as shown by the arrows.
Housing 210 and wall 211 prevent undesirable air leakage. Fan 221
moves the secondary air out of the heat exchanger in exhaust stream
203. Drains 222 may be included at the bottom of the housing to
remove excess water.
 Simple Solar Collector: FIG. 7 is a simple solar collector
for regenerating a desiccant solution. Liquid header, 230 trickles
desiccant liquid 238 over collector surface 231 which is tilted
at an angle to allow drainage through trough 234. Solar radiation
233 is transmitted through cover 235 and warms the desiccant liquid
238. The collector surface is preferably black and is backed by
thermal insulation 232 to maximize energy collection.
 The cover is preferably of a transparent material such as
polycarbonate, polyvinyl chloride (pvc), fluoropolymers (such as
Tedlar), acrylic, or other plastic. For rigid matericals, the cover
may be flat or corrugated. Flexible films such as Tedlar would normally
be held taught in a frame. Glass is another option for a cover material.
The selection of optimum collector material depends on the cost
and durability of the different materials. The duty is similar to
that for greenhouses, windows, skylights, etc. with temperatures
that are much lower than those for most other types of solar collectors.
 Incoming air 239 flows by natural convection between the
cover and the collector surface and absorbs moisture that evaporates
from the desiccant liquid. The leaving air 237 exits the collector
between end piece 236 and cover 235. The end piece is shaped so
as to prevent rain from entering the collector.
 The desiccant solution 238 preferably flows in a sheet that
wets the entire collector surface. One way of achieving this flow
is to use a screen, cloth, or other roughness on collector surface
231. Another option is to use a relatively large flow of liquid
to create a continuous film of liquid. A third option is to add
a detergent or other wetting agent to the solution to enhance wetting.
Combinations of these three alternatives are also possible.
 The orientation of the collector is preferably such that
the rays of midday summer sun is approximately normal to the collector
surface. For the most of the US this corresponds to a tilt of angle
of 5 to 30 degrees from horizontal towards the south. In tropical
areas the collector surface can be nearly horizontal with only a
few degrees of tilt to allow adequate drainage. In the Southern
Hemisphere the collector is preferably tilted to face north.
 The operation of this collector is quite simple. When solar
radiation is available to raise the collector surface to a temperature
that is sufficiently high, desiccant liquid is allowed to flow through
the collector. At other times no liquid would flow. A simple thermostat
that controls the circulating pump can accomplish this control.
 This collector has several advantages. First the temperatures
necessary to regenerate the desiccant liquid are quite low, in the
range of about 110 to 140 degrees Fahrenheit, which allows the use
of inexpensive materials such as plastic, wood, asphalt roofing
material, etc. Second the operation is very simple with no moving
parts. Third the collector can be mounted on an existing roof or
 While the preferred embodiment of the solar collector includes
a cover, the collector would also function without a cover. The
main advantages eliminating the cover are reduced cost and complexity.
The collector can, in fact, be as simple as a section of dark roof
or other surface with the addition of a system desiccant liquid
over the surface. The chief problem with operation without a cover
is that rain would tend to wash away any residual desiccant solution.
The resulting diluted desiccant would have to be discarded or else
it would dilute the solution in storage. Wind or leaves may also
carry desiccant solution away when no cover is present. Loss of
large quantities of desiccant solution is costly, may damage nearby
plants or metals, and may create unsightly salt deposits on surrounding
surfaces. A simple cover should greatly reduce or eliminate these
problems, but it is not absolutely necessary for operation.
 In dry climates an evaporation pond is an alternative to
a solar collector. A pond is an inexpensive way of regenerating
a desiccant solution. The chief problems are related to control
over the salt concentration. An extended rainy period can dilute
the solution excessively, while long periods of dry, sunny conditions
can result in crystallization. Another issue is the possibility
of high winds blowing droplets of desiccant solution onto surrounding
surfaces which may create problems with corrosion, plant damage,
 Solar Still with Automatic Shut-Off Feature: FIGS. 8a and
8b shows a solar still with an automatic shutdown feature that can
regenerate the desiccant liquid. FIG. 8a shows the still in normal
operation. Solar radiation 260 warms desiccant liquid 250. Water
vapor evaporates from the desiccant liquid and form condensate 252
on cover 251. The condensate trickles down the inside of the cover
and collects in troughs 253 and 254. An insulated tank 256 forms
the bottom and the sides of the collector and holds the desiccant
 Float 255 provides a simple control mechanism. As shown
in 7a, the desiccant solution is relatively dilute, which reduces
its density and causes the float to sink to the bottom of the tank
256. FIG. 8b shows a situation where the desiccant solution is quite
concentrated, which increases its density and causes the float to
rise to the top of the pool of desiccant liquid. Water evaporates
out of any remaining desiccant liquid on the surface of the float
and eventually creates a thin layer of salt crystals 257 which
helps to reflect solar energy and controls the temperature inside
the collector. The action of the float thus provides an automatic
shutdown feature that prevents excessive crystallization of the
 The float should be of nearly neutral buoyancy with respect
to the desiccant solution, so that the change in solution density
is enough determine whether the float rises or sinks. The float
materials should be resistant to high temperatures and compatible
with the desiccant solution. Foam glass, ceramics, high-temperature
plastics, and metals that are compatible with the desiccant are
likely choices. The float may be divided into smaller pieces to
 The sealed cover has the advantage of keeping ambient moisture
out of the desiccant during dark or cloudy periods. A cover without
a seal would allow free movement of humid air, which can add undesirable
moisture to the desiccant solution.
 High-Performance Solar Collector with Electric Back-Up:
FIG. 9 is a high-performance solar collector with electric backup
for use in regenerating the desiccant liquid. This collector provides
three stages of regeneration and would operate with peak temperatures
of around 200 to 240.degree. F. The three stages of regeneration
are arranged so the waste heat from higher-temperature stage drives
a lower temperature stage. Desiccant liquid 301 flows from bottom
header 304 over collector surface 300. Electric heater elements
302 are located just under the collector surface and provide an
auxiliary source of heat. Insulation 303 prevents excessive heat
loss through the back of the collector.
 The collector has three covers. The bottom cover 311 and
middle cover 312 fit tightly with frames 309 and 310 to minimize
air leakage. The top cover 313 has large gaps at each end, which
allow for air movement under the cover. The bottom and middle covers
311 and 312 are preferably of glass or other heat resistant material.
The top cover is preferably made of a tough plastic material to
minimize risks of hail damage or other hazards. The top cover experiences
much lower temperatures, so heat resistance is not an important
 Covers 311 312 and 313 transmit solar radiation 321 which
warms collector surface 300. At night or during cloudy periods,
electric heater elements 302 provide an auxiliary heat source for
warming the collector surface. The warm temperatures cause moisture
in the bottom stream of desiccant liquid 301 to evaporate. The water
vapor thus produced is moved by convection and/or diffusion to bottom
cover 311 where it condenses to form condensate 305. The condensate
flows down the undersurface of bottom cover and collects in a trough
formed by catch member 307 and frame 309. The condensate then flows
out of the collector and can be used in an evaporative heat exchanger,
as distilled drinking water, or discarded. Likewise a portion of
the desiccant liquid that collects at the bottom of the collector
can be returned to storage and the rest can be recirculated.
 The middle stream of desiccant liquid 317 flows from middle
header 314 over the top of the bottom cover 311. The heat transmitted
to the bottom cover from below evaporates moisture from the middle
stream of desiccant liquid 317. The moisture condenses on the bottom
surface of the middle cover 312 to form the middle condensate stream
306 which drains through the trough formed by catch piece 308.
 The top header 315 supplies the top desiccant liquid stream
316 that flows across the top surface of the middle cover 312. Moisture
that evaporates from the top desiccant liquid stream 316 is removed
by natural convection of air and does not normally condense on the
top cover 313. Entering air 323 flows through the collector and
receives the evaporating water vapor. End piece 320 prevents excessive
amounts of rain from entering the collector and air leaves the collector
as exhaust stream 322.
 While this figure shows an electric heater as the backup
heating system other heat sources are possible. Hot water, steam,
and direct heating with a fuel are also possible. The surface may
be heated directly or the desiccant liquid can be heated in a separate
heat exchanger. For small systems, a gas water heater may provide
the heat source. The selection of the heat source would be determined
by fuel cost and availability, installed cost and other factors.
If electric power is used it would preferably used at night to take
advantage of lower off-peak electric rates.
 The optimum collector temperatures depend on the desiccant
concentration, the ambient conditions, and other factors. For calcium
chloride a minimum temperature difference of about 30.degree. F.
is necessary to evaporate a desiccant solution and condense the
resulting water vapor. Assuming a temperature difference of about
40.degree. F. in each stage, temperature of the middle cover would
be about 120.degree. F., the bottom cover would be 160.degree. F.
and the collector surface would be 200.degree. F. If the peak temperature
is a problem, a two-stage system can be used instead but with a
performance penalty. Of course four or more stages are also possible,
but the collector temperature is normally limited to the boiling
point of the desiccant liquid which would be roughly 230.degree.
F. or somewhat higher.
 This collector requires a means for circulating the desiccant
liquid. Air-lift pumps or conventional pumps are possible alternatives.
The desiccant liquid would normally be recirculated several times
through the collector before returning to the storage tank. This
recirculation ensures sufficient movement of liquid for adequately
wetting the surfaces inside the collector without creating excessive
heat loss. A heat exchanger between liquid entering and leaving
the collector would further reduce heat losses, but this feature
is not required for operation of the system.
 The covers need to be tilted by roughly 10 degrees or more
to ensure proper drainage of condensate. Smaller angles could result
in excessive drainage of condensate back into the desiccant, which
would create a large performance penalty.
 The actual collector surface can be horizontal, which can
allow desiccant to pool inside the collector. This arrangement allows
the collector to also function as storage tank. Using a float as
a shut-off control as shown earlier would be desirable for this
system. This configuration may be especially desirable in tropical
areas where the sun's rays are nearly vertical during much of the
 Preferred Embodiment: FIG. 10 is a schematic drawing of
a complete solar air-conditioning system. Solar collector 400 concentrates
desiccant solution using heat input from solar radiation 424 or
auxiliary heat source as explained in the description of FIGS. 8A
and 8B. A tank 401 stores a large quantity of desiccant solution
402. The weak desiccant solution 420 leaves tank 401and flows through
the solar collector 400 and returns as a concentrated desiccant
 The amount of desiccant solution in the tank depends on
the size and efficiency of the system and the length of operation
required. A reasonable objective would be to achieve one to three
days of storage capability. For a system with a rated capacity of
12000 Btu/hr system with a 50% duty cycle over two days, corresponds
to storage requirement of 288000 Btu of cooling (24 ton-hours).
For a cooling COP of 1 and heat of vaporization of 1000 Btu/lbm
mean that this cooling requirement could be met by the ability to
absorb 288 lbm of water vapor. For calcium chloride solution with
a starting concentration of 50% and an ending concentration of 40%
CaCl.sub.2 by weight, requires two pounds of calcium chloride to
absorb one pound of water. This analysis means that 576 lbm of calcium
chloride is required to store the required cooling. For a 40% final
concentration, this corresponds to a tank capacity to handle 1440
pounds of solution or about 150 gallons.
 The energy storage density per unit volume is almost twice
that of ice and requires no special insulation. With proper sealing,
the tank can store this cooling capacity indefinitely with essentially
zero loss. The cost of calcium chloride is on the order of $0.20/lbm
so that the cost of the salt for the above example is a little over
$100. The cost of storage tank is similar. These costs work out
to be roughly $10 per ton-hour, which is roughly 10 to 20% of the
cost of conventional ice storage or cold-water storage. This storage
system thus has major cost and performance advantages compared to
other systems. This inexpensive, compact storage capability combined
with simple, efficient solar recharging is a tremendous improvement
over the prior art.
 Theory of Operation of the Coolers: FIG. 11 shows how temperatures
vary through a system for supplying outside air. The layout of the
modeled system is similar to that shown in FIG. 4. This system has
12 stages. Each stage has an air-to-air heat exchanger between primary
and secondary, a direct evaporative cooling section on the secondary
side, and a desiccant section on the primary side. In this analysis
only stages 1 2 and 4 have active desiccant sections. The mass
flow rate on the secondary air stream is approximately half of that
of the primary air stream for this analysis. The entering air conditions
are 95.degree. F. dry bulb and 75.degree. F. wet-bulb temperature,
which is a typical design condition for the eastern US.
 The system in FIG. 11 can be used to supply outside air
to laboratories or other applications that have a large outside
air requirement and limitations on heat recovery or other energy-saving
technologies. It has the advantage of conditioning outside air with
extremely high efficiency and requires no access to exhaust air.
Changes in the details of the design can give different supply-air
conditions as required for a particular application.
1TABLE 1 Temperature Enthalpy Location (degrees F.) (Btu/Ibm) Stage
Location Primary Air Stream: 500 95.0 38.6 1 Inlet 501 91.6 37.8
1 After heat exchanger 502 93.3 37.8 1 After desiccant 503 89.9
37.0 2 After heat exchanger 504 92.1 37.0 2 After desiccant 505
88.6 36.1 3 After heat exchanger 506 88.6 36.1 3 After desiccant
507 85.6 35.4 4 After heat exchanger 508 89.7 35.4 4 After desiccant
509 85.9 34.5 5 After heat exchanger 510 85.9 34.5 5 After desiccant
511 82.6 33.7 6 After heat exchanger 512 82.6 33.7 6 After desiccant
513 79.8 33.0 7 After heat exchanger 514 79.8 33.0 7 After desiccant
515 77.2 32.3 8 After heat exchanger 516 77.2 32.3 8 After desiccant
517 75.0 31.8 9 After heat exchanger 518 75.0 31.8 9 After desiccant
519 73.0 31.3 10 After heat exchanger 520 73.0 31.3 10 After desiccant
521 71.3 30.8 11 After heat exchanger 522 71.3 30.8 11 After desiccant
523 69.9 30.5 12 After heat exchanger 524 69.9 30.5 12 After desiccant
(supply air) Secondary Air Stream: 525 88.1 45.9 12 After direct
evaporative cooler 526 81.2 44.3 12 After heat exchanger 527 86.5
44.3 11 After direct evaporative cooler 528 79.7 42.7 11 After heat
exchanger 529 85.0 42.7 10 After direct evaporative cooler 530 78.0
41.0 10 After heat exchanger 531 82.6 41.0 9 After direct evaporative
cooler 532 76.7 39.6 9 After heat exchanger 533 82.2 39.6 8 After
direct evaporative cooler 534 74.7 37.8 8 After heat exchanger 535
79.4 37.8 7 After direct evaporative cooler 536 72.9 36.2 7 After
heat exchanger 537 77.0 36.2 6 After direct evaporative cooler 538
71.4 34.9 6 After heat exchanger 539 74.8 34.9 5 After direct evaporative
cooler 540 69.9 33.7 5 After heat exchanger 541 72.9 33.7 4 After
direct evaporative cooler 542 68.7 32.7 4 After heat exchanger 543
71.2 32.7 3 After direct evaporative cooler 544 67.5 31.8 3 After
heat exchanger 545 69.7 31.8 2 After direct evaporative cooler 546
66.6 31.0 2 After heat exchanger 547 68.6 31.0 1 After direct evaporative
cooler 548 66.0 30.4 1 After heat exchanger (exhaust)
2TABLE 2 Temperature Enthalpy Location (degrees F.) (Btu/Ibm) Stage
Location Primary Air Stream: 600 74.0 30.2 1 Inlet 601 73.6 30.1
1 After heat exchanger 602 81.1 30.1 1 After desiccant 603 78.6
29.5 2 After heat exchanger 604 80.7 29.5 2 After desiccant 605
77.9 28.8 3 After heat exchanger 606 77.9 28.8 3 After desiccant
607 75.5 28.2 4 After heat exchanger 608 78.4 28.2 4 After desiccant
609 75.4 27.5 5 After heat exchanger 610 75.4 27.5 5 After desiccant
611 72.6 26.9 6 After heat exchanger 612 72.6 26.9 6 After desiccant
613 70.2 26.3 7 After heat exchanger 614 70.2 26.3 7 After desiccant
615 68.0 25.7 8 After heat exchanger 616 68.0 25.7 8 After desiccant
617 65.9 25.2 9 After heat exchanger 618 65.9 25.2 9 After desiccant
619 64.1 24.8 10 After heat exchanger 620 64.1 24.8 10 After desiccant
621 62.4 24.3 11 After heat exchanger 622 62.4 24.3 11 After desiccant
623 60.9 24.0 12 After heat exchanger 624 60.9 24.0 12 After desiccant
(supply air) Secondary Air Stream: 625 73.2 36.0 12 After direct
evaporative cooler 626 72.5 35.8 12 After heat exchanger 627 76.1
35.8 11 After direct evaporative cooler 628 71.2 34.7 11 After heat
exchanger 629 75.1 34.7 10 After direct evaporative cooler 630 69.5
33.3 10 After heat exchanger 631 73.0 33.3 9 After direct evaporative
cooler 632 68.2 32.1 9 After heat exchanger 633 72.3 32.1 8 After
direct evaporative cooler 634 66.3 30.7 8 After heat exchanger 635
69.9 30.7 7 After direct evaporative cooler 636 64.5 29.4 7 After
heat exchanger 637 67.8 29.4 6 After direct evaporative cooler 638
62.9 28.2 6 After heat exchanger 639 65.8 28.2 5 After direct evaporative
cooler 640 61.4 27.2 5 After heat exchanger 641 63.9 27.2 4 After
direct evaporative cooler 642 60.0 26.2 4 After heat exchanger 643
62.3 26.2 3 After direct evaporative cooler 644 58.7 25.4 3 After
heat exchanger 645 60.7 25.4 2 After direct evaporative cooler 646
57.5 24.6 2 After heat exchanger 647 59.5 24.6 1 After direct evaporative
cooler 648 56.7 23.9 1 After heat exchanger (exhaust)
 FIG. 12 shows air temperatures for a system for cooling
return air. This system is similar to that of FIG. 11 except for
the different entering-air conditions. Table 2 describes each location.
This system is capable of providing a supply air temperature of
61.degree. F., which is sufficiently low to be compatible with most
conventional air-distribution systems. The supply air is at about
75% relative humidity, which is sufficiently low to prevent mold
growth in ducts and maintain comfortable space humidity.
 These supply-air conditions are illustrative of what is
possible with this system. Changes in the details of the design
can change the supply air conditions to whatever is required. For
example this system is suitable for use in an application which
uses a 65 to 70.degree. F. supply air temperature such as is described
in my co-pending application entitled "High Efficiency Air
Conditioning System with High Volume Air Distribution." Low
temperature air is also possible, but with reduced performance.
 A complete cooling system is a combination of the system
in FIGS. 10 and 11. The system in FIG. 12 exhausts approximately
half of the return air. The system in FIG. 11 supplies the make-up
air required to replace this exhaust air.
 High System Efficiency: The efficiency of these systems
is quite high. For these systems the coefficient of performance
(COP) is defined as cooling output divided by latent heat absorbed
by the desiccant. For the system in FIG. 11 the COP is approximately
2.1. For the combined system including ventilation load the efficiency
is about 1.5.
 Note that this configuration requires the introduction of
outside air. If the basis of comparison is system with no outside
air, then there should be no credit for the load associated with
cooling the outside air to the building conditions. On this basis
the system COP is approximately 0.8.
 If the desiccant is recharged using a solar collector, the
collector efficiency must be considered. For the three-stage collector
shown if FIG. 9 the waste heat from one stage is used to drive
the next, which theoretically can triple the output of concentrated
desiccant solution. In real life, the collector loss would reduce
the effect. Assuming a 50% collector loss, the total system efficiency
based on solar input to cooling output can exceed 2.0. This performance
is much better than what is possible with expensive absorption chillers
with high-performance solar collectors, which gives a system efficiency
of 0.2 to 0.5 at best. This efficiency advantage translates into
a massive reduction in collector cost and area required to drive
the new system compared to the prior art. Even with the use of low-cost
collectors such shown in FIG. 7 the new system has a massive efficiency
advantage compared with the prior art.
 This high performance allows the use electric resistance
as a back-up heat source. The efficiency with electric backup should
be higher than that for a solar input, since transmission and reflection
losses are not a factor. This means that system efficiency in the
range of 2 to 3 based on electric input is possible. This efficiency
is comparable to that of conventional electric air conditioners.
If combined with a suitable storage system, electric back up can
take advantage of inexpensive, off-peak electric rates. These rates
can be as much as a factor of 10 lower than peak rates. The combination
of solar input and off-peak rates can result in a massive reduction
in energy cost compared to conventional systems.
 Gas Turbine Inlet Cooler: FIG. 13 shows a system for cooling
inlet air to a gas turbine that uses heat from the turbine for regenerating
the desiccant solution. Ambient air 701 enters a regenerative desiccant
cooler 700 which cools the air. The cooler includes pumps for circulating
water and desiccant solution inside the cooler. The cooler receives
make up water 710 and used water 711 drains from the cooler. Exhaust
fan 714 draws the exhaust air 709 from the cooler and discharges
away from the turbine to prevent recirculation. The turbine inlet
air 702 leaves the cooler 700 and enters the gas turbine 713. Fan
715 adds ambient air 704 to the turbine exhaust air 703 to form
mixed air 705. The mixed air 705 enters regenerator 706 where it
evaporates water from the desiccant solution. The regenerator comprises
an extended surface that is wetted with desiccant liquid. It may
be made of materials similar to that used in direct evaporative
coolers. Mixing ambient air with turbine exhaust lowers the temperature
of air entering the regenerator, which allows the use of inexpensive
low-temperature materials. Pump 716 moves diluted desiccant solution
708 from the cooler 700 to the regenerator 706. The regenerator
can include a pump or circulating desiccant liquid inside the regenerator.
Concentrated desiccant 707 returns to the cooler from the regenerator.
Outlet air 712 exits from the regenerator.
 This system can increase turbine output power by roughly
20 percent at summer peak conditions. The capacity gas turbine declines
by about 0.4 percent per degree Fahrenheit. A 20% improvement in
capacity corresponds to cooling the inlet temperature is reduced
from 100 to 50.degree. F. at peak conditions. The system can also
control relative humidity to the turbine. Turbine efficiency improves
by roughly 0.1%/.degree. F. which corresponds to as much as 5% improvement
at peak conditions. Input power for fans and pumps needed to operate
the desiccant system is small and should not significantly effect
 Heat-Exchanger Details: FIGS. 14a, 14b, and 14c show an
alternate gas-to-gas heat exchanger design using paper and cardboard.
Sheets of paper 800 are supported between first spacers 801 and
second spacers 802. The spacers are preferably made of corrugated
cardboard or similar material. As shown in FIG. 14b, first spacers
801 are oriented to allow primary air stream 803 to flow the length
of the paper in a single pass. Second spacers 802 are set to form
multiple passes of secondary air stream 804. The whole heat exchanger
is coated with a material such as linseed oil, acrylic, wax, etc.
which serves as both an adhesive and a protective coating. While
this drawing shows a two-pass arrangement on the secondary side,
similar geometries can accommodate any number of passes. This heat
exchanger construction has many applications including exhaust-air
heat recovery in addition to use in evaporative and desiccant systems.
 Dehumidifer Embodiment: FIG. 15 shows another preferred
embodiment that is acts as a dehumidifier. Desiccant fluid 900 is
contained in an insulated container 904 and is heated by solar radiation
that is transmitted through a cover 902. Moisture evaporates from
the fluid and forms condensate 906 on the bottom side of the cover.
The condensate flows down the underside of the cover and collects
in a channel 908. The desiccant fluid moves by natural convection
through channels 910 and 912 to a tank 918. An air pump 914 blows
air through a tube 915 into the desiccant fluid 900 forming bubbles
916. The flow of air mixes the desiccant liquid in the tank. A fan
922 draws air over the desiccant fluid, which dehumidifies the air
stream. A baffle 920 directs the air to toward the surface of the
liquid. The baffle also acts to cut off air flow if the liquid level
gets too high, thus preventing and overflow of desiccant liquid.
A level switch may also be included to turn off the air pump and
fan at high liquid levels.
 This embodiment may be useful as a small dehumidification
system that can fit in a window. It may be especially useful for
bathrooms or basements in homes.
 FIG. 16--Preferred embodiment with liquid-to-liquid heat
exchange: FIG. 16 shows a preferred embodiment desiccant cooler
with a heat exchanger between desiccant liquid and water. The system
comprises a desiccant cooler 1008 a desiccant tank 1061 and a solar
collector 1090. A fan 1022 draws ambient air and moves an air stream
1024 through desiccant-gas heat and mass exchanger 1012 that cools
and dehumidifies the air. Air stream 1024 moves through an evaporative
cooler to form a supply air stream 1026 which is enters a conditioned
space 1050. A fan 1052 moves air from the conditioned space as an
air stream 1028 which then moves through a water-gas heat and mass
exchanger 1014. The water-gas heat and mass exchanger humidifies
and adds thermal energy to the air stream to create an exhaust air
stream 1030 which is discharged to the ambient. It also acts to
cool water 1046 which flows through the exchanger.
 There are two liquid loops that supply these two heat and
mass exchangers. A desiccant loop includes a pump 1036 which draws
liquid desiccant 1018 from a reservoir 1019 and supplies it at the
top of the desiccant-gas heat and mass exchanger 1012 as a drops
of water 1038 that wets media that forms the exchanger. The air
and desiccant are preferably in a counterflow configuration with
the desiccant flowing down while the air flows upward. Desiccant
collects in a drain pan 1032 and then flows through a liquid-to-liquid
heat exchanger 1010. The heat exchanger 1010 cools the desiccant,
which then flows back to reservoir 1019.
 The water loop is formed by a pump 1020 which draws warm
Water 1042 from liquid-to-liquid heat exchanger 1010 which preferably
has a counterflow arrangement. Water 1046 is distributed at the
top of the water-to-gas heat and mass exchanger 1014 and accumulates
in a drain pan 1040 and returns to the heat exchanger 1010 as cooled
water 1044. As with the desiccant side, the water and air are preferably
arranged in a counterflow configuration to maximize thermal performance.
Make-up water 1049 enters through a valve 1051 to replace blow-down
water 1047 which drains from the system and water that evaporates.
The valve 1051 may be a float valve or other device to maintain
an approximately constant volume of water in the water loop.
 The solar collector 1090 comprises a cover 1068 that transmits
solar radiation 1078 to warm a pool of desiccant liquid 1070. The
pool of desiccant liquid is preferably is contained by a liner 1082
which is supported by a layer of support material 1072 which sits
on top of a roof 1076 or other flat surface. The support material
is preferably stone or ceramic granular material such as pea-sized
gravel, pearlite, sand, or polystyrene foam beads. Closed-cell foam
in another option. The liner is preferably made of a black plastic
material such polyvinyl chloride, polyethylene, rubber such as is
used in pools liners or landfill liners. It preferably includes
a cloth or foam underlay for improved life and reduced heat transfer.
These liners are typically about 20 to 40 mils thick and are designed
to withstand years of ultraviolet radiation from the sun. A frame
1074 supports the cover 1068. The frame may be made of pressure-treated
lumber or other material suitable for outdoor use. Ambient air 1080
enters the collector through holes in the frame 1074 and flow through
the collector to remove moisture evaporated from the desiccant and
leaves as an exhaust stream 1082. The walls formed by the liner
1082 slope slightly to allow liquid to drain to a pump 1066.
 A desiccant storage tank 1061 contains a quantity of desiccant
1060 sufficient for at least two hours of operation of the cooler
1008. The preferred quantity depends on the climate conditions,
but would normally be at least sufficient to allow operation from
late afternoon to the next morning when solar input is limited or
not existent. A pump 1064 moves desiccant liquid to the solar collector
during period when solar energy is available and the desiccant solution
needs to be further concentrated. A pump 1062 moves liquid to the
cooler 1008 as required to maintain a proper desiccant concentration
in the cooler. An optional heat exchanger may be included between
the desiccant entering the reservoir 1019 and leaving the reservoir
so as to reduce thermal losses associated with the fluid transfer.
 The system preferably includes a controller 1092 that controls
operations of the fans and pumps. The controller is in communication
with a solar sensor 1094 and liquid-level sensor 1093. The solar
sensor may comprise a black thermistor or other temperature sensor
that is exposed to solar radiation. The liquid-level sensor is preferably
a simple liquid-level switch. The controller also receives input
from a desiccant concentration sensor 1095 that preferably comprises
a float switch that closes when the desiccant density reaches a
predetermined value. Other options include more sophisticated sensors
such as density sensor or electrical conductivity measurements or
a simple liquid level sensor.
 The controller uses the input from these two sensor to determine
when operate the solar collector. When the solar sensor 1094 senses
a sufficiently high temperature (about 100 to 130 F) and the concentration
sensor shows that the desiccant is sufficiently dilute, the controller
turns on pump 1064 to move a quantity of desiccant to the collector
1090. Once the liquid level reaches a predetermined limit, the liquid-level
sensor 1093 communicates to the controller and the control turns
off the pump 1064. The collector preferably has sufficient liquid
holding capacity, that in the event of a failure of the liquid level
sensor would not result in overflow of desiccant liquid. The controller
activates the pump 1066 and pump 1062 to periodically move concentrated
desiccant back to the tank 1061 which accumulates concentrated
liquid desiccant. A heat exchanger may be included between the desiccant
entering and leaving the collector to improve thermal performance.
 During extended periods of very dry weather, it may be necessary
to include a means for adding water to the desiccant to prevent
crystallization of salt. This situation may occur in desert climates
where ambient dewpoint temperatures are lower than about 55 F, but
sensible cooling is still required. The water addition is preferably
accomplished by diverting circulating a portion of the exhaust stream
1030 into the air entering fan 1022 which raises the dewpoint of
air entering the desiccant-gas heat and mass exchanger 1012. Alternatively,
the evaporative cooler 1016 may be operated without the rest of
the cooler 1008 during periods of sufficiently low humidity without
diverting exhaust air.
 A humidity sensor 1096 and a temperature sensor 1098 which
are located in the conditioned space 1050 provide input to control
the operation of the cooler 1008. The humidity sensor may be a humidistat
and the temperature sensor may be a thermostat. If the temperature
or humidity exceeds predetermined limits, then the controller turns
on the cooler 1008. If the humidity is sufficiently low but the
space is too warm, then the controller may activate the evaporative
cooler 1016 to provide a lower temperature for the supply air stream
1026 otherwise the evaporative cooler is normally off.
 For good performance of the cooler 1008 it is necessary
to maintain proper flow for the fluids. In a typical application,
the flow rate of the air stream 1028 should be close to that for
the supply air stream 1026. In addition the flows of the water and
the desiccant liquid should be adjusted to maintain close the same
temperature change across the heat exchanger 1010. The liquid temperature
change should also be close the airside temperature changes for
good performance of the heat exchangers. For simplicity, these adjustments
are preferably made manually and set at a design value. Alternatively,
for optimum performance, the adjustments can be made continuously
with an automatic controller that receives input from appropriate
 FIG. 17--Chiller embodiment: FIG. 17 shows an embodiment
that is suitable for producing chilled water for air conditioning.
A cooler 1 100 comprises a desiccant-gas heat and mass exchanger
1104 and water-gas heat and mass exchanger 1102 in an enclosure
11 12. The enclosure is preferably gas-tight and capable of withstanding
atmospheric pressure when a partial vacuum is created inside. A
fan 1110 circulates gas between the two exchangers. A vacuum pump
1114 draws gas 1116 from the space and discharges it to the atmosphere
 The operating pressure inside the enclosure 1112 is preferably
about 1 to 5 psia. Gas filling the enclosure is mixture a mixture
of air and water vapor. Lowering the atmospheric pressure has several
advantages. First it greatly improves the heat- and mass-transfer
coefficients. These coefficients are approximately inversely proportional
to the partial pressure of the air, which means they get very large
as the pressure of the gas mixture approaches the vapor pressure
of the water. A second advantage is that the lower pressure reduces
the thermal losses associated with circulating the gas between the
two exchangers. A third advantage is that the lower pressure reduces
the fan energy required to circulate the gas.
 While air is preferred for simplicity and low-cost, low-molecular-weight
gas such as hydrogen or helium may be used instead to improve heat
transfer, in which case gas should be recovered from the exit of
the vacuum pump. While operation at below atmospheric pressure is
preferred, the system can work at atmospheric pressure, but with
a large performance penalty.
 Dry gas 1120 leaves the top of the desiccant-gas exchanger
1104 and enters the bottom of the water-gas exchanger 1102. The
exchangers are arranged in a counterflow configuration with liquid
entering at the top and gas entering as the bottom. The temperature
of water drops as it flows through the water-gas exchanger 1102
and the exit temperature approaches the wet-bulb temperature of
the gas entering the exchanger. Humid gas 1122 leaves the top of
the water-gas exchanger 1102 and enters the bottom of the desiccant-gas
 The cooler include provisions for changing the working liquids.
A pump 1136 removes a small portion of the circulating water to
prevent excessive accumulation of salts. Make-up water 1134 replaces
water withdrawn by the pump along with water evaporated in the water-gas
exchanger 1102. Likewise a pump 1140 removes a quantity of the circulating
desiccant solution, which is replaced by concentrated desiccant
1142 to ensure a proper concentration of desiccant.
 Pumped fluid transfers heat outside of the cooler. Cooled
water 1126 is circulated by pump 1124 through a coil 1128. A fan
1130 moves air over the coil 1128 to cool a conditioned space 1132.
A pump 1148 moves desiccant 1126 through a heat exchanger that is
cooled by water from a cooling tower 1144. A pump 1146 circulates
water through heat exchanger 1150 to the cooling tower 1144.
 The cooler 100 differs from an conventional absorption chiller
in that it is designed to handle a gas-vapor mixture. In conventional
chillers, the systems are designed to operate with extremely low
levels of non-condensable gases. Any appreciable quantity of non-condensable
gas creates large heat transfer penalties because there is no fan
for moving the gas across the heat exchange surfaces. By comparison
the present invention uses fan and an extended heat/mass exchange
surface to allow operation with a large amount of non-condensable.
This ability to handle large quantities of non-condensable gas greatly
improves the ability to use storage of desiccant and regeneration
of desiccant at atmospheric pressure without special concerns about
contact with air.
 The present cooler is designed to take advantage of a temperature
glide inherent in a gas-vapor mixture. In a conventional absorption
chiller, the vapor pressure of the evaporator and the absorber is
a single value. In contrast, air or other non-condensable gas allows
the vapor pressure to vary, while maintaining close to a constant
total pressure. This difference increases the available temperature
lift from the desiccant by an amount that roughly corresponds to
the temperature change of the cooling water across the absorber.
 The optimum performance of the cooler 1100 occurs when the
temperature change of each fluid is approximately the same value.
This setup minimizes the temperature drop through each heat exchanger
and improves the temperature lift capability.
 For calcium chloride the maximum available temperature lift
(entering desiccant temperature minus leaving chilled water temperature)
is about 25 F. For temperatures lifts greater than this amount a
two-stage system or a different desiccant is required.
 For climates with relatively low design wet-bulb temperatures
(such as those in California), the cooler should be able to provide
sensible cooling in a single-stage configuration. For example, for
a design wet-bulb temperature of 70 F, a leaving cooling-tower water
temperature of 75 F is reasonable (5 F approach temperature). A
temperature of 77 F for the desiccant leaving heat exchanger 1148
should allow a water temperature to the coil of about 60 F, which
should allow a supply air temperature of about 68 F. This setup
is especially suitable for use with the air conditioning system
described in U.S. Pat. Nos. 6405543 and 6185943 which use high-temperature
air distribution system with separate dehumidification.
 FIGS. 18 and 19--Rotating Embodiment: FIGS. 18 and 19 show
an alternate embodiment of a cooler with rotating heat exchangers,
which preferably comprises multiple stages. FIG. 18 shows a detail
of a single stage. A pipe 1200 encloses a first direct-contact heat
exchanger 1202. The interior of the pipe also contains a liquid,
preferably the desiccant 1210 with end pieces 1206 that prevent
leakage from the ends.
 A second direct-contact heat exchanger 1212 is located around
the outside of pipe 1202. Tubing 1208 is wrapped around the second
direct contact heat exchanger and is connect at each end through
fittings 1204 and 1205 to the inside of pipe 1202. The bottom portion
of the tubing and the second direct-contact heat exchanger sit in
pool of liquid, which is preferably water, which is located below
liquid level line 1214.
 The whole assembly rotates as a unit, which provides a means
for circulating the liquids. Liquid desiccant 1210 enters into tube
1208 through fitting 1204. The rotation and the force of gravity
move the desiccant through the tubing and it eventually returns
to the inside of pipe 1200 through fitting 1205. The turning action
also submerges portions of the first direct-contact heat exchanger
in desiccant 1210 and portions of the second direct-contact heat
exchanger 1212 into water located below liquid line 1214. These
setup allows circulation of liquid for heat and mass transfer without
the use of a pump.
 FIG. 19 shows a multi-stage assembly for this cooler. A
first, second, and third stage 1244 1246 and 1248 are all connected
together a rotate as a unit. Each stage has a geometry that is similar
to that in FIG. 18. Pans 1241 are connected together with tubes
1242 and are filled with water. Make up water 1254 enters the pan
for the first stage 1244 and blow-down water 1252 exits the pan
for the third stage. Rollers 1242 support the stages and allow them
to rotate freely. A motor 1260 that is connected to the rotating
stages by a shaft 1262 turns the assembly. Ambient air 1270 is drawn
by fan 1250 and moves through the desiccant side of the stages.
Supply air 1272 exits the first stage and cools a conditioned space
1264. Return air 1274 flow through the waterside of the assembly
and exits as exhaust air 1276. While three stages are shown the
cooler may use twelve or more stages, depending on the design requirements.
 Alternate thermal-energy input: While solar energy input
is preferred in many applications, there are situations where solar
is not practical because of space limitations, climate, or other
factors. In those situations the preferred embodiment uses an alternate
source of thermal energy. While natural gas or other fuel is one
alternative, availability and/or cost may limit its use. Another
alternative is to use waste heat from a conventional vapor-compressor
or absorption refrigeration system.
 Resistance heat from off-peak electricity is yet another
alternative. The system can regenerate the desiccant at night or
on weekends during periods of low electric prices for use during
periods of high electric prices.
 For these systems without solar input, no collector is required.
Instead a regenerator, preferably with multiple stages of regeneration,
may be included. Various arrangements similar to those found in
the prior art for distillation of sea water, absorption chillers,
etc. are possible. For a regenerator that is limited to atmospheric
pressure and peak temperatures of about 200 to 250 F, three stages
of regeneration should be possible with calcium chloride. This temperature
and pressure is achievable using relatively inexpensive and corrosion-resistant
material such as plastic and ceramic in the construction of the
 For more stages, higher pressures and temperatures are required.
Theoretically ten stages or more of regeneration are possible. More
stages of regeneration increase cost and complexity of the regenerator,
but improve the efficiency of the system. The optimum design depends
on material costs, pressure-vessel code considerations, cost of
the thermal input, temperature limits of available materials, and