Hybrid air conditioning system subsystem integration

A hybrid air-conditioning system having an absorption refrigeration subsystem for sensible heat loads and a cooperating liquid desiccant dehumidification subsystem for latent heat loads incorporates additional apparatus for effecting the efficient transfer of available heat between the subsystems to improve the total system Coefficient of Performance (C.O.P.) over a wide range of latent heat load to sensible heat load proportions. Available heat recovery and system internal load generation are utilized in high and low latent load ratio regions, and modulated heat transfer using an additional series absorption refrigeration cycle loop is optionally utilized in the mid-latent heat load ratio region.

FIELD OF THE INVENTION 
The present invention relates generally to air conditioning, and 
particularly concerns improved apparatus and methods for reducing the 
temperature and relative humidity of air circulated within an enclosed 
space such as a building interior. 
More specifically, the invention concerns a hybrid-type of air conditioning 
system wherein sensible heat loads are handled by an absorption 
refrigeration subsystem, wherein latent heat loads are handled by a 
cooperating liquid desiccant dehumidification subsystem, and wherein 
variations in systems loads and load proportioning are handled by novel 
subsystem integration apparatus and methods to improve total system 
thermodynamic efficiency performance. 
BACKGROUND OF THE INVENTION 
Numerous applications of desiccant dehumidification to the conditioning of 
air are known in the prior art. Meckler U.S. Pat. Nos. 3,401,530 and 
3,488,971, and Griffiths U.S. Pat. No. 4,164,125, for instance, utilize a 
solid desiccant for the application. Similarly, Meckler U.S. Pat. No. 
4,011,731, and Turner U.S. Pat. No. 4,171,620 teach the use of a desiccant 
in the conditioning of air but emphasize the use of liquid desiccant 
materials. Also, Meckler's U.S Pat. No. 3,102,399 suggests a building air 
conditioning system wherein make-up ventilation air is subjected to liquid 
desiccant dehumidification in a two-stage dehumidification process to 
improve total system performance efficiency but is forced to use a 
two-stage dehumidification process. 
Meckler et al. U.S. Pat. No. 4,171,624 teaches the use of thermal 
compressor means to regenerate or concentrate a dilute desiccant solution. 
Meckler also, in U.S. Pat. No. 4,222,244 for example, teaches the use of 
solar energy in desiccant regeneration for an air conditioning system. See 
also Meckler U.S. Pat. No. 4,577,471 in the regard. 
Griffiths U.S. Pat. No. 4,259,849 also teaches the use of heat obtained 
from the condenser of a conventional vapor compression refrigeration 
system for effecting liquid desiccant regeneration. 
Meckler U.S. Pat. No. 3,247,679 discloses an engine-driven vapor 
compression refrigeration subsystem in a comfort conditioning system that 
also utilizes a liquid desiccant dehumidification subsystem. Meckler's 
U.S. Pat. No. 3,153,914 teaches air conditioning with a liquid desiccant 
dehumidification dehumidifier but without supplemental refrigeration. 
U.S. Pat. No. 2,981,078 discloses air cooling dehumidification and 
dehumidification using a hydroscopic agent and a rotating foraminous disk 
partially immersed in the agent. Supplemental absorption or mechanical 
refrigeration is not suggested. 
Taylor U.S. Pat. No. 2,355,828 discloses an earlier combined refrigeration 
and dehumidification air conditioning system. 
Mattern, et al. U.S. Pat. No. 2,262,954, discloses an air dehumidification 
system with controls to prevent desiccant crystallization during liquid 
desiccant regeneration. Richardson U.S. Pat. No. 20,257,204 also teaches 
liquid desiccant regeneration in a manner that improves the reclamation of 
waste heat. 
For other variations of air conditioning systems employing liquid desiccant 
solutions for dehumidification of air, see U.S. Pat. Nos. 4,635,446, 
4,691,530, and 4,723,417, all issued in the name of Meckler. 
SUMMARY OF THE INVENTION 
A building air conditioning system configured in accordance with the 
present invention is comprised of a refrigeration subsystem and a 
cooperating liquid desiccant dehumidification subsystem. The refrigeration 
subsystem, in one embodiment of the invention, is preferably an absorption 
chiller fueled by a natural gas energy source. Such refrigeration 
subsystem provides available heat to the liquid desiccant dehumidification 
subsystem for the purpose of effecting or assisting in effecting liquid 
desiccant regeneration (dilute desiccant solution concentration) in the 
latter subsystem without penalizing the efficiency of the refrigeration 
subsystem. Also, the refrigeration subsystem is provided in the invention 
for the purpose of effecting air temperature variation and control by 
handling the system sensible heat load associated with air circulated 
within the building air conditioned, enclosed space. 
The liquid desiccant dehumidification subsystem is provided in the 
invention for the purpose of effecting conditioned air relative humidity 
variation and control by handling the total system latent heat load 
associated with the air circulated within the building air conditioned, 
enclosed space. In the preferred embodiment of the present invention, a 
building air-conditioned, enclosed space is or may be continuously or very 
nearly continuously provided with fractional ventilation air from outside 
the enclosed space, and relative humidity control is effected by 
processing that ventilation air fraction through the liquid desiccant 
dehumidification subsystem to accomplish moisture removal from the 
processed air. The processed fractional ventilation air is then combined 
(mixed) with air recirculated from the enclosed space and the resulting 
mixed air is lowered in temperature by the air conditioning system 
refrigeration subsystem. 
To achieve an improved coefficient of performance for the total system, and 
also to achieve economic advantages by way of reduced equipment 
acquisition costs and by way of reduced operating fuel or energy costs, I 
utilize available heat from the refrigeration subsystem to effect, at 
least in part, either increased liquid desiccant regeneration in the 
liquid desiccant dehumidification subsystem portion of the total air 
conditioning system or increased refrigerant production for the absorption 
refrigeration subsystem potion of the total system. Such available heat 
may sometimes be characterized as reject heat. Additionally, particular 
cooling tower and air-to-air heat exchanger components may be 
advantageously incorporated into such liquid desiccant dehumidification 
subsystem to further improve total air conditioning system performance. 
When integrating the liquid desiccant dehumidification subsystem to 
configure the total hybrid air conditioning system, I recover heat from 
the refrigeration subsystem condenser and/or absorber elements through 
various system operation steps and/or system added apparatus elements to 
improve total system thermodynamic efficiency performance in manners which 
accommodate varying system sensible heat loads and latent heat loads, both 
as to their individual magnitudes and relative proportioning. Such may be 
accomplished without the necessity of providing system supplementary heat, 
and in certain cases is accomplished by switching the utilization of 
available heat, for instance, from desiccant regeneration to simple heat 
rejection in response to a reduced latent heat load. Other invention cases 
involve the utilization of modulator assemblies interposed cooperatively 
between the system refrigeration subsystem and the system desiccant 
regenerator unit and operable to utilize either refrigeration subsystem 
absorber reject heat recovery or condenser reject heat recovery, 
preferably with a degree of heat recovery modulation, to significantly 
improve total system performance in a manner which is somewhat more 
complex from an apparatus standpoint. Thus, the present invention 
basically includes the utilization of control apparatus and/or operating 
parameter modification, either individually or in combination, to shift 
"reject" heat flow progressively to either of two thermodynamically useful 
system functions--increased desiccant regeneration or increased 
refrigerant production--by means of novel thermodynamic cycles. Such 
utilization involves the recovery of heat values from fixed heat flows in 
the refrigeration subsystem condenser, absorber, or combined condenser and 
absorber elements, from modulated heat flows in the refrigeration 
subsystem condenser, absorber, or combined condenser and absorber 
elements, and from combined fixed and modulated heat flows in such 
refrigeration subsystem elements and element combinations, all in relation 
to the system dehumidification subsystem liquid desiccant regeneration 
apparatus and methods. 
The foregoing and other advantages of the invention will become apparent 
from the following disclosure in which a preferred embodiment of the 
invention is described in detail and illustrated in the accompanying 
drawings. It is contemplated that variations and structural features and 
arrangement of parts may appear to the person skilled in the art, without 
departing from the scope of sacrificing any of the advantages of the 
invention which is delineated in the included claims.

DETAILED DESCRIPTION OF THE INVENTION 
Hybrid air conditioning systems of the type generally combining an 
absorption refrigeration subsystem with a liquid desiccant 
dehumidification subsystem are previously known, and one such system 10 is 
illustrated by the functional block diagram of FIG. 1 and corresponds 
essentially to the type of hybrid air conditioning system disclosed in 
allowed U.S. patent application Ser. No. 07,302,428, filed Jan. 27, 1989, 
and issued Mar. 6, 1990 as U.S. Pat. No. 4,905,479. 
Such system includes in part, an air distribution subsystem 11 that 
recirculates air returned from within an enclosed building space for 
reconditioning which typically involves air cooling and a lowering of air 
relative humidity. Cooling is accomplished in system 10 by absorption 
chiller refrigeration subsystem 12 and dehumidification or relative 
humidity control is accomplished by liquid desiccant dehumidification 
subsystem 13. Subsystem 13 may process fractional ventilation air received 
from the ambient atmosphere outside the system and also a controlled 
portion of the return air as diverted from air distribution subsystem 11. 
Such ventilation air, after processing in the dehumidifier unit portion 15 
of subsystem 13 for dehumidification, is flowed back to air distribution 
subsystem 11 for mixing with the remaining return air and for subsequent 
cooling by chilled water received from refrigeration subsystem 12. 
Refrigeration subsystem 12 is conventional in construction and typically 
includes, inter alia, an evaporator heat exchanger component, a condenser 
heat exchanger component, and an absorber heat exchanger component. Also, 
such typical absorption chiller subsystem includes an energy or heat 
source, such as a natural gas burner (not shown in the drawings). Further, 
refrigeration subsystem 12 may be provided in the form of a double-effect 
subsystem. Subsystem 12 functions to desorb, condense, evaporate, and 
re-absorb refrigerant from an absorption refrigerant pair such as a 
lithium bromide and water solution. The refrigerant passed through the 
absorption refrigeration subsystem evaporator heat exchanger cools or 
chills water that is circulated separately to and through distribution 
subsystem 11 to effect temperature changes in the combined recirculated 
and ventilation airflows and thus handle the building sensible heat load. 
The FIG. 1 cooling tower subsystem 21 is conventional and cooperates with 
the absorption refrigeration subsystem 12 through a subsystem in the 
refrigeration absorber heat exchanger component and with the liquid 
desiccant dehumidification subsystem 13 through a heat exchanger in the 
dehumidification subsystem dehumidifier unit 15. Also, the liquid 
desiccant dehumidification subsystem 13 cooperates with absorption 
refrigeration subsystem 12 through the flow of relatively dilute desiccant 
solution to and through the (dilute) subsystem 12 condenser exchanger 
component. The desiccant dehumidification subsystem 13 is principally 
comprised of dehumidifier unit 15 and desiccant regenerator unit 22, and 
the desiccant solution utilized in subsystem 13 is typically an aqueous 
solution of lithium chloride, lithium bromide, ethylene glycol, etc. 
A liquid desiccant such as a relatively concentrated lithium chloride 
solution is flowed in the prior hybrid system from the collector of 
desiccant regenerator unit 22 to a sprayer in subsystem dehumidifier unit 
15. The concentrated (or "strong") desiccant solution is sprayed in 
dehumidifier unit 15 into the stream of ventilation air received from air 
distribution subsystem 11 and reduces the water content of that 
ventilation air to the desired reduced level by absorption of the "excess" 
moisture into the desiccant solution thus causing the solution to become a 
diluted (or "weak") desiccant solution. The dilute desiccant solution is 
flowed in heat exchange relation to the refrigeration subsystem condenser 
heat exchanger component for heating and afterwards the heated desiccant 
relatively dilute solution is sprayed into a desiccant packed regeneration 
bed 39 situated in the ductwork of regenerator unit 22. 
As schematically illustrated in FIG. 1, hybrid system 10 may optionally 
also include an auxiliary heat source such as auxiliary gas burner 24 
which is selectively operable to heat a medium in heat exchanger 37 and to 
transfer such heat to the relatively dilute desiccant solution flowed from 
the condenser heat exchanger component of subsystem 12 to desiccant 
regenerator unit 22 for regeneration. Such optional auxiliary burner 24 is 
normally actuated only when the latent heat load portion of the system 
total load is relatively much larger than the system sensible heat load 
portion. 
FIG. 2 provides a diagram for a conventional double effect absorption 
refrigeration cycle wherein refrigerant solution vapor pressure (P) is 
plotted as the ordinate and vapor temperature (T) and decreasing vapor 
concentration (X) are plotted as abscissas. Referring to FIG. 2, points A 
and B bound the absorption process in a conventional single-effect cycle 
in which concentrated refrigerant solution at the concentration defined by 
Point A is diluted to its equilibrium exit condition, point B, by 
refrigerant vaporized at conditions represented by point C. The suppressed 
temperature of the evaporation, point C, accomplishes the cooling while 
the temperatures of the absorption process, A-B, are sufficient to reject 
heat to the ambient. Dilute refrigerant solution at B is pumped to a 
pressure sufficiently higher than the pressure at E to allow flow through 
the recuperative heat exchange path between points B and E. The dilute 
solution heated to approach point E is fed to a desorber where it is 
concentrated to the condition of point F by the external heat input that 
drives the system. This concentrated solution is cooled from point F to A 
as the dilute solution fed to the desorber is heated to E'. The 
equilibrium point E is used for convenient reference and that temperature 
is not usually attained by the recuperative preheat (heat exchange). 
The vapor generated by the desorption process EF is condensed by cooling 
tower water at a condition represented by point D. The liquid refrigerant 
at D (H.sup.2 O) is reduced to the pressure at C as it is flowed into the 
evaporator where its evaporation accomplishes the desired cooling. 
The P,T,x cycle of FIG. 2 utilizing a "double-effect" involves thermal 
cycling from points H to B, B to K, K to L, L to M, M to I, and I to H. 
Points H and B bound the absorption process in a conventional 
double-effect cycle in which concentrated refrigerant solution at the 
concentration defined by point H is diluted to its equilibrium exit 
condition, point B, by refrigerant vaporized at conditions represented by 
point C. Dilute refrigerant solution at B is pumped to a pressure 
sufficiently higher than the pressure at K to allow flow through a 
two-step recuperative heat exchange path between points B and K. The 
dilute solution heated to approach point E is further heated to point K 
and fed to a high pressure desorber where it is partially concentrated to 
the condition of point L by the external heat input that drives the 
system. This partially diluted solution is cooled from point L to point M 
as the concentrated solution fed to the high pressure desorber is heated 
from E' to K'. The equilibrium point K is used for convenient reference 
and that temperature is not usually attained by the recuperative pre-heat 
(heat exchange). 
The pure refrigerant vapor (e.g., 100% water) released from the high 
pressure desorption process, K-L, is condensed at temperature J which is 
higher than the temperature between points M and I so that this 
condensation energy can further concentrate the solution in the low 
pressure desorption process between points M and I. The vapor released 
during the low pressure desorption (M to I) is condensed at temperature D. 
The low pressure condenser D rejects heat to the ambient whereas the high 
pressure condenser J, rejects heat to the low pressure desorber, M-I. The 
additional concentration of solution in the low pressure desorber is 
driven by heat released within the cycle (high pressure condensation at J) 
and is referred to as the "second effect". The fully concentrated solution 
at I is cooled from I to H (approximately) as it heats the dilute solution 
from B to E. 
FIG. 3A is an energy flow diagram for the single-effect absorption 
refrigeration cycle of FIG. 2. Energy from the combustion products enters 
the solution desorber with some energy being lost as combustion products 
exhausted to the flue. Refrigerant vapor generated in the desorber is 
condensed into useable refrigerant liquid. As this refrigerant is 
condensed its condensation energy is used to regenerate the liquid 
desiccant circulated through the liquid desiccant dehumidification 
subsystem. The energy released from the absorber is conventionally 
rejected to the cooling tower water. 
FIG. 3B is an energy flow diagram for the basic FIG. 2 double-effect cycle. 
Energy from the combustion products enters the first-effect solution 
desorber (generator) with some energy lost as the combustion products are 
exhausted to the flue. Refrigerant vapor is generated in the first-effect 
desorber and, in the process of being condensed into usable refrigerant 
liquid, drives additional refrigerant from solution in the second-effect 
desorber. As this additional refrigerant is condensed into usable liquid 
refrigerant its condensation energy is used to regenerate the liquid 
desiccant. 
Although usually combined in a single evaporator, the two liquid 
refrigerant flows are each shown to be capable of the separate cooling 
effects shown on FIG. 3. Similarly, two absorption processes in which the 
evaporated refrigerant is (re)absorbed into solution are shown, even 
though the two processes occur simultaneously in one absorber. The energy 
released from the absorber is conventionally rejected to the cooling tower 
water. The evaporation processes and the absorption processes are both 
shown as two separate processes for thermodynamically clarity. 
In the liquid desiccant dehumidifier, the desiccant solution is 
sufficiently concentrated that it extracts water vapor directly from the 
air passing through it. The energy released by the change of state of the 
water (from vapor to a portion of liquid solution) heats the desiccant 
solution which is cooled by cooling tower water. The desiccant 
dehumidification process has thus converted the latent load into a 
sensible thermal load that can be rejected from the system by the "free" 
evaporative cooling provided by the cooling tower. 
The dehumidification capacity, however, is not defined exclusively by the 
cooling tower capacity. The water absorbed into the liquid desiccant 
solution must be removed from the desiccant so the concentrated liquid 
desiccant solution can be recycled to the dehumidifier. In this 
"regeneration" process, the desiccant is heated so that water essentially 
boils (evaporates) from the solution when contact with a flow of outside 
air occurs. The hotter the desiccant, the more water that can be driven 
off. But with the elevated desiccant temperature some of the regeneration 
energy is lost to sensible heating of both the ambient transport air and 
the desiccant solution itself. In practice, only slightly more than 70% of 
the regeneration energy actually removes water vapor from the desiccant 
solution so that a like amount can be added in the dehumidification 
process. No more latent load capability can be created than that 
associated with desiccant regeneration. FIG. 3A illustrates the limited 
latent load (46% of the fuel input to the absorber) that can be obtained 
from the condenser reject heat in the known basic hybrid air conditioning 
system. FIG. 4 shows the performance characteristics for the known basic 
single-effect and double-effect cycles through a plot of the pertinent 
overall system Coefficient of Performance (C.O.P.) as a function of the 
latent load fraction--the ratio of the latent load to the total load. At 
the extreme left ordinate, the operating condition with no latent load, 
the C.O.P. is 1.05, that assumed for the absorption chiller alone. As the 
latent load fraction increases, progressively more of the condenser reject 
heat is used for desiccant regeneration and the C.O.P. increases. This 
C.O.P. increase is accompanied by an increase in total load capacity equal 
to the amount of latent load. At the "Match Point", all the condenser 
reject heat is used for desiccant regeneration. At higher latent load 
ratios, the auxiliary burner 24 supplies the additional regeneration 
energy and the composite C.O.P. decreases. A lower C.O.P. limit of about 
0.7, that of the liquid desiccant system by itself, is reached when the 
chiller is not fired because there is no sensible load, and all the 
desiccant regeneration energy is supplied by the system auxiliary burner. 
Since the accumulated ton-hours of system operation over an entire season 
would be dominated by operation relatively close to the Match Point for 
most buildings, the seasonal C.O.P. is expected to be quite high, as shown 
illustratively on FIG. 4. In short, the basic hybrid air conditioning 
system has the potential to improve the operation of any given chiller by 
a least 25 percent, providing its condensing temperature can be increased 
by about 20.degree. C. (36.degree. F.) without exceeding structural 
pressure limits of the system. 
In reviewing the performance characteristics depicted by FIG. 4, the most 
significant deficiencies exist at latent load ratios above the "Match 
Point". It is therefore desirable to both increase the system C.O.P. in 
the mid-range of latent load ratios and to avoid the substantial fall-off 
of performance caused by the utilization of auxiliary or supplemental 
energy (heat). This can be achieved in accordance with this invention in a 
number of different manners of system operation with correspondingly 
different system apparatus configurations. Basically such methods of 
operation and apparatus variations are keyed to the more efficient 
utilization of potentially available reject heat from different system 
sources such as multiple-effect absorption refrigeration cycles and 
refrigeration subsystem solution recrystallization control additives and 
the control of that reject heat utilization for desiccant regeneration 
purposes. Table 1 below introduces, by different designations, the 
different inventive approaches described and claimed herein. 
"Reject heat" from the system condenser or absorber, or from both condenser 
and absorber, can be raised in temperature in different ways and be and 
used in varying proportions so it can be useful for desiccant regeneration 
purposes. The first class of desired system rejection heat modifications 
is identified by the first horizontal row in Table 1. In this Class I, the 
quantity of reject energy entering the desiccant system is "fixed". Such 
available desiccant regeneration heat may exceed the actual latent load 
regeneration heat requirement so that varying degrees of regeneration 
efficiency will be experienced to keep the system in balance. In essence, 
it is a characteristic of this class of hybrid system that the potentially 
recoverable "excess" regeneration energy may be switched from desiccant 
regeneration to simple rejection in response to a particular latent load. 
It is also characteristic of this "Fixed" class of system modification that 
the other rejection heat source component (absorber or condenser) operates 
normally unless specified otherwise. In a Cf system, for example, the 
condenser element becomes elevated in temperature due to reduced sensible 
heat load and is coupled to the desiccant regenerator unit while the 
absorber element rejects its heat to the cooling tower at normal 
temperatures. 
The known prior art hybrid air conditioning systems if utilizing available 
heat for desiccant regeneration are limited to this Cf classification. In 
the Af system, for example, the solution concentrations are defined so 
that the absorber element will operate at an elevated temperature and be 
coupled to the desiccant regenerator unit while the condenser element 
rejects its heat to the cooling tower at normal temperatures. As shown in 
Table 1, the third option in classification I involves operation of both 
the absorber and the condenser at temperatures sufficiently elevated that 
increased desiccant regeneration is possible. 
TABLE 1 
______________________________________ 
Heat Recovery Classifications 
Source of Recovered Heat 
Condenser 
Absorber Both 
______________________________________ 
I Fixed Heat Flow 
Cf Af CfAf 
II Modulated Heat Flow 
Cm Am CmAm 
III Combined Fixed & 
CfAm CmAf (CmAm) 
Modulated 
______________________________________ 
Also identified in Table 1 are a second (II) and a third (III) class of 
improved hybrid systems modification. The Class II modifications come from 
a desire to utilize the absorption system reject heat (made 
thermodynamically "more available" by raising its temperature) even when 
maximum desiccant regeneration is not needed. This would improve operation 
to the left of the "Match Point" on FIG. 4. In this Class II, the reject 
heat from the designated component is sufficiently elevated that it can be 
used for desiccant regeneration and/or for the production of refrigerant 
in an additional absorption cycle. The "Modulated" title refers to the 
progressive shifting of heat flow between two thermodynamically useful 
functions--desiccant regeneration and refrigerant production--in a 
separate ("Modulated") cycle. In the modulated configurations of this 
invention, desiccant regeneration improvement is developed as between 
elements of two absorption refrigeration cycles interconnected in a series 
or tandem relation. Just as two absorption cycles can be integrated so 
that a driving cycle can be considered a "first effect" and a second 
absorption cycle can be considered a "second effect" since it is arranged 
to be powered or driven by heat rejected from the driving cycle, "reject 
heat" or better "available heat" in this invention is used in two 
functions--for increased driving of the second absorption refrigeration 
cycle, for increased regeneration of liquid desiccant, or for any 
intermediate combination. It is from this continuously variable split of 
the useful available heat between the two functions (refrigerant 
production or desiccant regeneration) that the "modulated" classification 
is derived. Consistent with the classifications of the Fixed class (Class 
I) systems in Table 1, the following definitions are made for "Modulated" 
(Class II) systems: 
In the Cm system, heat from the condenser in the driving absorption cycle 
is elevated in temperature so that it can be used for desiccant 
regeneration and/or for refrigerant production. The absorber heat from the 
driving cycle is rejected at normal temperatures to the cooling tower 
water. Heat from the modulated absorption cycle absorber and condenser 
also is rejected at normal temperatures directly to the cooling tower 
water. 
In the Am system, heat from the absorber in the driving cycle is elevated 
in temperature for either desiccant regeneration and/or refrigerant 
production. Heat from the driving cycle condenser, from the modulated 
cycle condenser, and from the modulated cycle absorber goes directly to 
the cooling tower water. The absorption fluid in the driving cycle must 
run at high concentrations to achieve the required temperature elevation. 
In the CmAm system, heat from both the absorber and the condenser of the 
driving cycle is elevated so that it can either regenerate desiccant 
and/or generate refrigerant in the modulated cycle whose condenser and 
absorber reject heat directly to the cooling tower water and operate at 
normal temperatures. 
The third class (III) of modified hybrid air conditioning system listed in 
Table 1 represents the logical combinations of the two previous classes. 
The example that will be developed is a CmAf system in which increased 
temperature condenser heat from the driving cycle is modulated between 
desiccant regeneration and refrigerant production while the heat from the 
driving cycle absorber is only fed to the desiccant regeneration system. 
Inherent in the modulated subsystem is an additional set of absorption 
cycle components. The desorber is part of the modulator unit and, where 
the concentrations are different, a separate absorber is needed. 
Condensing and evaporation functions can be integrated into the driving 
cycle components. 
No definition of the individual absorption cycle complexity is established 
in Table 1, but it can be assumed that, unless defined otherwise, both the 
input (driving) absorption cycle and the modulated absorption cycle are 
single-effect cycles. Subscripts 2 or 3 appended to such designations 
hereinafter will be utilized to indicate double-effect and triple-effect 
absorption refrigeration cycles when applicable. One method of identifying 
or designating the potentially different subsets and combinations of this 
invention utilizes the generalization of a Cz(r)Az(n) hybrid air 
conditioning system wherein: z is specified as "m" for modulated or "f" 
for fixed heat recovery; (r) is the number of condenser-coupled effects in 
the modulated cycle; and (n) is the number of condenser-coupled effects in 
the driving cycle. The absence of (r) or (n) implies a single effect. 
FIGS. 5 and 6 of the drawings illustrate with block diagram depiction the 
modifications to the known hybrid air conditioning system to result in two 
different Class II (Modulated) hybrid air conditioning systems of Table 1. 
The block diagram system 30 of FIG. 5 is for a Table 1 Cm-type modulated 
hybrid air conditioning system; the block diagram system 40 of FIG. 6 is 
for a Table 1 Am-type system. As defined above, the driving absorption 
refrigeration system 12 in FIGS. 5 and 6 could have (n) condenser-coupled 
effects where n equal 1,2,3, etc. 
System 30 also includes a modulator unit 31 which operatively couples 
absorption refrigeration subsystem 12 to the desiccant regenerator unit 22 
of desiccant dehumidification subsystem 13. In the case of system 40, 
however, the included modulator unit is designated as 41 even though 
generally has the same basic physical tube-in-tube heat exchanger 
configuration as modulator unit 31. Modulator unit 31 incorporates, as 
element 32, the condenser heat transfer surface of the driving cycle 
condenser element of absorption refrigeration subsystem 12. Modulator unit 
41, on the other hand, incorporates, as element 42, the absorber heat 
transfer surface of the driving cycle absorber element of absorption 
refrigeration subsystem 12. Although modulator unit heat exchangers 32 and 
42 of FIGS. 7 and 8 are illustrated as having a single tube within a tube 
assembly, in practice and for the reason of heat exchange surface 
adequacy, such units are preferred to each be constructed of multiple tube 
within a tube assemblies interconnected by a common vapor header and a 
common lower liquid collection sump. Also, the illustrated spiral fluted 
tube is of a commercially available configuration as is the illustrated 
integrally spined tube. Such spines may project either upwardly and 
outwardly or downwardly and outwardly without a significant degradation of 
performance. Further, the individual members of each tube pair are 
essentially in contact or near contact with each other to define a spiral 
passageway for desiccant solution that is restrictive to enhance heat 
transfer over the spiral flow path. 
Referring to FIGS. 7 and 9, modulator unit 31 for a condenser-modulated 
hybrid air conditioning system 30 has a heat exchanger assembly 32 that is 
provided with multiple openings 33 through 39. Relatively hot refrigerant 
vapor from the refrigeration subsystem desorber assembly (FIG. 9, 101) is 
flowed into inlet opening 33, condenses on the tubular absorption 
refrigeration subsystem condenser heat transfer surface 51, collects in 
the heat exchanger assembly 32 sump, and is afterwards flowed through 
outlet opening 34 to the evaporator absorber assembly (FIG. 9, 102) of the 
absorption refrigeration subsystem. Similarly, relatively dilute modulator 
unit refrigeration solution, which like the refrigeration solution of 
subsystem 12 may be a LiBr and H.sub.2 O solution, is flowed into inlet 
opening 37, wets and is vaporized on the spirally-fluted tubular modulator 
unit desorber heat transfer surface 52, and its water vapor constituent is 
afterwards flowed through outlet opening 38 to the modulator unit 
condenser assembly (FIG. 9, 53). Concentrated modulator unit refrigeration 
solution collects at the bottom of the annular chamber defined by heat 
transfer surface 51 and the outer shell of assembly 32, and is flowed 
through outlet opening 39 to the heat exchanger assembly 55 and afterwards 
to the modulator unit absorber/evaporator assembly 54 where it absorbs the 
refrigerant evaporated in the cooperating element in assembly 54 Heat 
transfer between tubular heat transfer surfaces 51 and 52 is effected in 
part by the controlled and variably restricted flow of dehumidification 
subsystem relatively dilute desiccant solution through inlet opening 35, 
between tubular elements 51 and 52, and afterwards through outlet opening 
36 to the system desiccant regenerator unit 22. 
Referring additionally to FIG. 9, modulator unit 31 also includes a 
condenser assembly designated 53, an evaporator/absorber assembly 54, a 
further heat exchanger 55, and pump means 56 which circulates relatively 
dilute modulator unit refrigeration solution from evaporator/absorber 54, 
through heat exchanger 55, and to inlet opening 37 of heat exchanger 
assembly 32. Various line means are specifically shown in FIG. 9 to 
illustrate the flow paths of relatively hot refrigeration subsystem 
refrigeration solution vapor and relatively cool refrigeration liquid 
(H.sub.2 O) to and from openings 33 and 34, of relatively cool and 
relatively hot dilute dehumidification subsystem desiccant solution to and 
from openings 35 and 36, and modulator unit refrigerant solution flows of 
dilute refrigerant to solution and refrigerant vapor and concentrated 
refrigerant solution from heat exchanger assembly openings 37, 38, and 39. 
Other line connections illustrating the flow paths of cooling tower water 
to and from subsystem 21 and chilled water to and from air distribution 
subsystem 11 are also shown. 
As shown in FIGS. 8 and 10, modulator unit 41 has a configuration and a 
heat exchanger assembly 42 that differs from the modulator unit details of 
FIG. 7 and 9. Heat exchanger assembly 42 of the absorber modulated hybrid 
air conditioning system is provided with multiple openings 43 through 50. 
Relatively dilute modulator unit refrigerant solution is flowed from the 
modulator unit evaporator/absorber assembly (FIG. 10, 64) through inlet 
opening 43 for concentration at the desorber heat transfer surface 60 and 
is desorbed vapor constituent (H.sub.2 O) afterwards is flowed through 
outlet opening 44 to the condenser assembly 63 of modulator unit 41. A 
more concentrated refrigerant solution collects in the lowest sump of heat 
exchanger assembly 42, and is flowed through outlet opening 50, through 
counterflow heat exchanger 65, and to evaporator/absorber assembly 64 to 
absorb evaporated refrigerant (vapor). Conversely, relatively concentrated 
refrigeration solution from the absorption refrigeration subsystem (FIGS. 
10, 12) desorber assembly (FIG. 10, 201) is flowed into inlet opening 47, 
and simultaneously refrigeration solution vapor from the absorption 
refrigeration subsystem evaporator element (FIG. 10, 202) is flowed 
through inlet opening 48. The concentrated refrigeration solution flowed 
through inlet opening 47 flows over and wets the exterior surface of 
spined tube 61 and absorbs the refrigerant vapor received through opening 
48 to produce relatively dilute refrigerant solution that is collected in 
the lower most extreme of the annular absorber chamber formed between tube 
61 and the outer shell of heat exchanger assembly 42. The collected 
relatively dilute refrigeration subsystem refrigerant solution is 
afterwards flowed through outlet opening 49, through heat exchanger 203, 
and in to the system desorber assembly 201 (FIG. 10). Heat transfer 
between tubular desorber heat transfer surface 60 and tubular absorber 
heat transfer surface 61 is effected in part by the controlled and 
variably restricted flow of dehumidification subsystem relatively dilute 
desiccant solution through inlet opening 45, between tubular elements 60 
and 61, and afterwards through outlet opening 46 to the system 40 
desiccant regenerator unit 22. The heat generated by the absorption at 
surface 61 is controllably split between sensible heating of the desiccant 
solution entering at inlet opening 45 and the desorption of refrigerant 
(H.sup.2 O) from the relatively dilute refrigeration solution entering at 
inlet opening 43. 
Referring additionally to FIG. 10, modulator unit 41 also includes a 
condenser assembly designated 63, an evaporator/absorber assembly 64, a 
further heat exchanger 65, and a pump means 66 which circulates relatively 
dilute modulator unit refrigeration solution from evaporator/absorber 
assembly 64, through heat exchanger 65, and to inlet opening 43 of heat 
exchanger assembly 42. Various line means are additionally shown in FIG. 
10 to illustrate the flow paths of modulator unit refrigeration solution 
relatively dilute liquid and vapor constituents to and from openings 43 
and 44, of absorption refrigeration subsystem refrigeration solution 
concentrated liquid, vapor, and dilute liquid constituents to inlet 
openings 47 and 48 and from outlet opening 49, and also dehumidification 
subsystem relatively dilute desiccant solution to and from openings 45 and 
46. Other line connections illustrating flow paths of cooling tower water 
to and from subsystem 21 in system 40 and chilled water to and from air 
distribution subsystem 11 in system 40 are also illustrated in FIG. 10. 
Modulator units 31 and 41 each include a valve, 57 or 67, which responds to 
a system "Humidistat" sensor and which controls or throttles the flow of 
dehumidification subsystem dilute desiccant solution to desiccant 
regenerator unit 22 for concentration in response to detected changes in 
the actual proportion of system sensible and latent heat loads. Also shown 
in the drawings are conventional absorption refrigeration subsystem 
expansion valves 105, 205 and also circulation pumps 106 and 206, which 
accomplish the flow of relatively dilute refrigeration solution in their 
respective absorption refrigeration subsystems. 
From a study of the delineated different hybrid air conditioning systems it 
becomes apparent that the primary cooling coils in the air distribution 
subsystem that received chilled water from evaporator 102 or 202 need not 
be cold enough to control humidity but it is important that the chilled 
water temperature be kept above the dewpoint of the processed air so that 
they will remain dry and allow the desiccant dehumidification subsystem to 
control air humidity level without interference. This means that an 
evaporator temperature of 50.degree. to 55.degree. F. should normally 
prove acceptable if the building air dewpoint is not less than 58.degree. 
F. A number of different evaporator refrigeration cycles can be matched to 
the modulator unit to avoid chilled water coil moisture condensation and 
the Coefficient of Performance characteristics of the different cycles are 
illustrated in FIG. 11 with their respective absorption refrigeration 
cycle appropriately indicated. 
For the Cm type subsystem an auxiliary burner (FIG. 1, 24) will be required 
if the system latent load factor exceeds about 40%. For the Cm(2) type 
subsystem, the maximum pressure for the double-effect absorption 
refrigeration cycle can be held to less than approximately one atmosphere 
gage if properly designed. In the Am cycle solution concentration is 
increased so that heat will be provided to the modulator unit and any 
required refrigeration solution formulation changes or solution additive 
changes are moderate and achievable. Similar solution concentration and 
additive adjustments are in order for the Af(n) subsystem cycles although 
the degree of adjustment required is less for Af(n) cycles than for Am(n) 
cycles. 
The performance plots CmAm and Cm(2)Am(2) relate to dual-coupled subsystem 
single-effect and double-effect cycles wherein available heat is taken 
simultaneously from the refrigeration subsystem condenser and from the 
refrigeration subsystem absorber thereby increasing system Coefficient of 
Performance over the single-coupled subsystem cycles (Cm(1) or Am(1) for 
example). Cycle maximum pressures for the latter dual-coupled system are 
well above any applicable one atmospheric gage pressure code restriction. 
In FIG. 12 I provide a P,T,x diagram for a representative Cm absorption 
refrigeration subsystem single-effect refrigeration cycle. When no 
desiccant regeneration is required (0% latent heat load), the 
single-effect driving absorption refrigeration cycle loop is from A to B 
to E to F to A with refrigerant-saturation conditions at corresponding 
points C and D. (See explanation above for FIG. 2, single-effect cycle). 
The single-effect modulated absorption cycle loop is from A to B to E' to 
F' to A when no heat is diverted to the desiccant loop for desiccant 
regeneration. It is, of course, logical that this P,T,x diagram is 
identical with a parallel-flow double-effect absorption cycle. The 
condensation process at D is hot enough to heat desiccant solution for 
regeneration whenever that is desired. When modulation occurs in the 
modulator unit to achieve an increased degree of desiccant regeneration, 
the single-effect driving cycle does not change but the concentration span 
between E' and F' shrinks since less heat is available for refrigerant 
production. The absorbers for the driving cycles and the modulator cycle 
can be merged into one unit physically since the leaving concentration at 
B is common. Other double-effect cycle arrangements, such as the series 
flow arrangement illustrated in FIG. 2, can be combined into a Cm type of 
hybrid air conditioning system. 
Similarly, the P,T,x diagram of FIG. 13 illustrates the single-effect 
absorption refrigeration cycle characteristics of an absorber-coupled 
modulator unit arrangement (Am(1) system). When no desiccant regeneration 
is required, the absorption refrigeration cycle loop is from A to B to E 
to F to A with refrigerant saturation conditions at points C and D. If no 
available heat is transferred from refrigerant production to desiccant 
regeneration, the modulated refrigeration cycle loop becomes A' to B' to 
E' to F' to A' with the same refrigerant saturation points C and D as the 
heat from absorption process A to B drives the modulated cycle desorption 
process from E' to F'. As this heat is diverted to the desiccant 
regeneration, the concentration span from E' to F' shrinks. 
FIGS. 14 and 15 are P,T,x and C.O.P. plots for a representative combined 
condenser-coupled and absorber-coupled system in which only available 
condenser heat is modulated. Such system is denoted a CmAf system, the 
single-effect characteristic being implied for both the driving and 
modulator unit cycles. As in many other cases, when absorber pressure 
levels can be raised, solution crystallization concerns are reduced. In 
the FIG. 14 plot, the refrigeration driving cycle is from A to B to C to D 
to A with refrigerant saturation at E and F. The modulator unit absorption 
cycle loop extends from G to H to I to J to G with the modulator units' 
condenser refrigerant solution condition being at point K. The temperature 
at point F is sufficiently elevated that it can provide heat for the 
desorption process I to J, but the solution temperatures in the absorber 
(A to B) are elevated only enough for desiccant regeneration and are not 
at temperatures high enough to transfer heat to the desorption from I to 
J. 
FIG. 15 shows the representative performance of the CmAf cycle as the 
latent heat load varies over the full range. At the zero latent load 
condition the C.O.P. is that of a double-effect absorption machine (e.g., 
C.O.P. of 1.0). At lower latent load fractions, no desiccant flow is sent 
to the modulator unit and all available heat from the driving cycle 
condenser is used to drive the modulated tandem absorption refrigeration 
cycle. The driving cycle absorber provides progressively more heat for 
desiccant regeneration as the latent load increases. In the middle latent 
load range, all the heat from the driving cycle desorber is used for 
desiccant regeneration, and the C.O.P. is relatively constant because it 
is perceived that desiccant regeneration derived from condenser heat at F 
is as thermodynamically advantageous as driving the tandem absorption 
cycle. The descending portion of the FIG. 15 C.O.P. curve at high latent 
load fractions does not involve use of an auxiliary heat source for 
desiccant regeneration. Instead, some of the refrigeration subsystem 
evaporative cooling capacity is used to cool the dehumidification process, 
thus lowering the temperature requirements for desiccant regeneration when 
no system sensible load exists, the cooling of the dehumidification 
process becomes the only sensible load imposed on the refrigeration 
subsystem evaporator, and this heat is in effect added to the input from 
the auxiliary burner to provide an enhanced amount of heat for desiccant 
regeneration. FIG. 15 shows the impact of this "heat pumping" effect on 
system C.O.P. at normal desiccant regeneration efficiencies. The chiller 
cooling of the dehumidification process will allow more dilute desiccant 
concentration to be used which in turn will improve regeneration 
efficiency and therefore improve total system C.O.P.s in the high latent 
load fraction region. 
A hybrid air conditioning system with the refrigeration subsystem 
evaporator sized for maximum sensible heat load will operate at part load 
virtually all the time. Because humidity control is not compromised at 
part load, improved comfort control is a bonus. 
Further system operating advantages may be realized if the disclosed two 
separate refrigeration solution loops which function in tandem are merged 
into one loop with a wide solution concentration range and the absorber 
heat used only for desiccant regeneration. This simple merger can be 
accomplished if the driving effect cycle becomes a double-effect cycle 
with the refrigeration subsystem absorber element operating over a broader 
concentration range and without modulated condenser recovery. The energy 
flow diagram of FIG. 16 confirms the desirability of this approach which 
is an implementation of an Af(2) type hybrid system as defined in Table 1. 
In an Af(2) type hybrid system the maximum available heat recovered from 
the driving cycle combined first and second-effect absorber elements very 
nearly equals the amount of heat recovered from the separate absorber and 
condenser elements in a CmAf type hybrid system (FIGS. 14 and 15), but 
none of the regeneration heat is heat diverted from refrigerant 
production. FIG. 16 shows the improved C.O.P. over the CmAf system to be 
limited in the mid-range. Where this mid-range dominates the cooling 
season, this is an effective alternative to the CmAf system. 
The functional block diagram and system schematic illustration for such 
Af(n) type hybrid system are essentially similar to like illustrations for 
the Cf or Cf(2) type hybrid system correspondingly illustrated in the 
aforesaid U.S. patent application Ser. No. 07/302,428 except that the 
available heat for desiccant regeneration is taken by heat exchange or 
heat transfer from the absorption refrigeration subsystem absorber rather 
than from the absorption refrigeration subsystem condenser. Depending on 
system permissible maximum operating pressure, the refrigeration 
double-effect cycle may take the form of a series flow double-effect cycle 
(FIG. 18) or a parallel flow double effect cycle (FIG. 19). The FIG. 18 
cycle is pressure (line J-K-L) can be lower than the peak pressure of the 
series cycle of FIG. 18. 
It is also possible to modify the Af(2) type system absorption 
refrigeration cycle in the manner indicated by the P,T,x diagram of FIG. 
19 to further reduce the maximum system operating pressure. Such requires 
the operation of an additional refrigeration solution circulation pump at 
solution temperatures above normal (e.g., 200.degree. F.) to feed partly 
concentrated solution through the heat exchanger--F to K and to the 
desorber K to L. Advantageously, the desorber pressure at L is fully 
available to force flow concentrated refrigeration solution back through 
the solution heat exchangers to the subsystem absorber element. Such an 
Af(2) system appears capable of operating at a peak pressure less than 
atmospheric, an advantage in some areas of the world where restrictive 
operating codes apply. For a relatively simple "spin-down" procedure to 
avoid solution crystallization after the shut-down of the subsystem 
desorber element driving burner it is only necessary to continue operating 
the circulation pumps through system cool-down to acceptable solution 
concentrations. Only modest quantities of anti-crystallization additives 
should be needed in this refrigeration subsystem absorption cycle 
approach. 
The desiccant solution flow control loop 300 schematically illustrated in 
FIG. 21 may be provided in the foregoing hybrid air conditioning systems 
of different designations to effect convenient system adjustment to load 
variations which require either increasing the recovery of available heat 
or increasing the cooling of desiccant solution by heat transfer to 
chilled water cooled by the refrigeration subsystem evaporator. Such 
increases respectively increase the system Coefficient of Performance in 
the regions of proportionately low and proportionately high system latent 
heat loads. For instance, the increased C.O.P. obtained in the Cm Am, 
Am(2), and CmAm(2) cycles of FIG. 11 over the remaining cycles wherein the 
C.O.P. equals approximately 0.7 may be achieved by the utilization of the 
FIG. 21 arrangement to cool the desiccant with chilled water, thus 
recovering heat for regeneration and eliminating the need of an auxiliary 
heat source such as a supplementary burner assembly. This is a critically 
advantageous trade-off since periods of excessive latent load are usually 
accompanied by low sensible loads. The total system capacity has the same 
shape as the C.O.P. plots in these examples. The high latent load factor 
regions (right-hand portion) of FIGS. 15 and 17 show this same improvement 
obtained by the utilization of the FIG. 21 arrangement. 
Referring to FIG. 22, relatively concentrated desiccant solution is flowed 
by pump means 304 from the sump 305 of desiccant regenerator unit 22 for 
subsequent cooling and spraying from nozzles 306 of desiccant dehumidifier 
unit 15. Heat exchangers 307, 311 and 312 and control valves 314 through 
316, and 318 are provided in the connecting lines of control loop 300 in 
the manner shown for selective utilization and actuation for acceptance of 
the latent heat load and the cooling of desiccant solution when latent 
heat loads are relatively low (e.g., 0% to 25-50%). 
With respect to the disclosed valving, valve 314 is responsive to the fluid 
level in sump 302 and is closed when further air dehumidification within 
dehumidifier unit 15 is not possible. Thus, a relatively dilute desiccant 
solution will continue to be circulated by pump means 301. When air 
dehumidification is required, valve 315 is opened in response to the 
signal from its' humidistat and adds concentrated desiccant to the flow 
sprayed from nozzle 306. This extracts moisture from the air and causes 
the level of fluid in sump 302 to rise so valve 314 is opened to flow 
dilute desiccant solution to regenerator unit 22 through heat exchangers 
307 through 310. During periods of excessive latent load, insufficient 
desiccant regeneration will have occurred so that insufficient 
dehumidification will occur with valve 315 wide open. Continued 
dehumidification demand will actuate valve 318, providing a second stage 
of dehumidification control. Opening valve 318 admits a flow of chilled 
water from the refrigeration subsystem 11 to heat exchanger 312 to lower 
temperatures than were possible by the cooling tower 12 in heat exchanger 
311 not only increases the dehumidification capability of the given 
desiccant solution but it increases the load on the chiller which 
increases the heat available for desiccant regeneration. Valve 316 is 
responsive to the temperature of the concentrated solution leaving heat 
exchanger 308 and is normally closed when system latent heat loads are 
proportionately high. The refrigeration subsystem 12 absorber element 
provides available heat to the flowed or recirculated desiccant solution 
through heat exchanger 309 if the system is designated (defined) as an Am 
or an Af system. Similarly, heat exchanger 310, which may optionally be 
provided in those Cm or Cf designated systems which have refrigeration 
cycles developing adequate high condenser temperatures for desiccant 
heating, will reject its heat to the flowed desiccant solution. For Am and 
Cm systems the heat is rejected directly to the desiccant (see FIGS. 7 and 
8) so heat exchangers 309 and 310 should be considered as schematic. In 
addition, these systems by their nature only reject heat to the desiccant 
when dehumidification is needed. However, for Af and Cf type systems, the 
reject heat is delivered to the desiccant regeneration. If the dilution of 
the desiccant in the dehumidifier unit is less than the concentration 
caused by the excessive reject heat, the temperature, the temperature 
controlling valve 316 will increase. Opening valve 316 admits cool water 
from cooling tower subsystem 21 so that the desiccant is cooled in heat 
exchanger 308 so that its regeneration will match the dehumidification. 
Closing of valve 316 accompanies the rise of C.O.P. from 0% latent load 
conditions such as displayed on the plots of FIGS. 4, 15, and 17. As 
indicated previously, the incorporation of desiccant solution control loop 
300 into a hybrid air conditioning system is particularly advantageous 
when the system otherwise has a low Coefficient of Performance (e.g., as 
low as 0.7) under 100% latent load conditions. 
It has been understood that although the present invention has been 
specifically disclosed with the preferred embodiments and examples, 
modifications and variations of the concepts herein disclosed may be 
resorted to by those skilled in the art. Such modifications and 
variations, for example flattening the sensor tip or altering the sensing 
element geometry, are considered to be within the scope of the invention 
and the appended claims.