Refrigeration – Atmosphere and sorbent contacting type
Reexamination Certificate
1999-10-08
2001-06-19
Buiz, Michael (Department: 3744)
Refrigeration
Atmosphere and sorbent contacting type
C062S094000
Reexamination Certificate
active
06247323
ABSTRACT:
TECHNICAL FIELD
This invention relates to air conditioning systems, and relates, in particular, to an air conditioning system that can operate continually an air dehumidifying process by a desiccant and regeneration of the desiccant by heat pump.
BACKGROUND ART
FIG. 5
shows a system based on the conventional technology disclosed in a U.S. Pat. No. 4,430,864, which is comprised by: process air passage A; regeneration air passage B; two desiccant beds
103
A,
103
B; a heat pump for regeneration of desiccant and cooling of process air. The heat pump uses heat exchangers
220
,
210
embedded in the two desiccant beds
103
A,
103
B as high and low temperature heat sources respectively, in which one desiccant bed performs dehumidifying by passing process air, and the other desiccant bed performs regeneration of desiccant beds by passing regeneration air. After air conditioning is carried out for a specific time interval, four-way switching valves
105
,
106
are operated to perform reverse processes in respective desiccant beds by flowing regeneration air and process air in the opposite desiccant beds.
In the conventional technology described above, high and low heat sources of the heat pump and each desiccant are integrated in each unit, and, an amount of heat equivalent to the cooling effect &Dgr;Q, is totally loaded on the heat pump (vapor compression refrigeration cycle). That is, cooling effect cannot exceed the capability of the heat pump (vapor compression refrigeration cycle) used. Therefore, there is no benefit resulting from making the system complex.
Therefore, to resolve such problems, it is possible to consider a system, such as the one shown in
FIG. 6
, to heat the regeneration air by placing a high temperature source
220
in the regeneration air passage B, and placing a low temperature air source
240
in the process air passage A to cool the process air, as well as to provide a heat exchanger
104
for exchanging sensible heat between the post-desiccant process air and pre-desiccant regeneration air. In this case, the desiccant
103
uses a desiccant wheel which rotates so as to straddle the process air passage A and the regeneration air passage B.
This system can provide cooling effect (&Dgr;Q), which is a sum of the cooling effect produced by the heat pump and the cooling effect produced by sensible heat exchange performed between process air and regeneration air, as shown in the psychrometric chart presented in
FIG. 7
, thus producing a system of a more compact design and capable of generating a higher cooling effect than that produced by the system shown in FIG.
5
.
In such a heat pump
200
, it is necessary to provide a high-temperature heat source with a temperature of over 65° C. for desiccant regeneration, and a low-temperature heat source with a temperature of about 10° C. for cooling process air. A vapor compression type cooling process for a refrigerant HFC134a is shown in a Mollier diagram shown in
FIG. 8
, and the temperature rise is 55° C., and the pressure ratio and compressor power are closer to the heat pump in a conventional air conditioning system based on refrigerant HCFC22. Therefore, there is a possibility of constructing a heat pump using a compressor for HCFC22 for desiccant regeneration in air conditioning systems.
However, in a system of such a configuration, if a conventional type of so-called plate-fin-coil heat exchanger of a single assembly unit, as shown in
FIG. 6
, is used as a high-temperature heat source heat exchanger, in which refrigerant flows through multiply branched refrigerant passages in a cross flow relation to the air, and the refrigerant and the air exchange heat in a state of disorderly temperature distribution, heat energy retained by the refrigerant cannot be transferred adequately to the air.
A relation between the temperature changes and enthalpy changes for the refrigerant and the regeneration air at the known heat exchanger is illustrated in FIG.
9
. As shown in
FIG. 9
, when refrigerant of the heat pump and regeneration air exchange heat, enthalpy changes for the refrigerant and regeneration air are equal due to heat balance. In the heat exchange process, refrigerant loses its enthalpy in a sensible heat change process from a superheated vapor state at the exit of the compressor until it starts condensation, and loses its enthalpy in a condensation process through a latent heat change process while maintaining its temperature constant, and further loses its enthalpy in a sensible heat change process from a saturated liquid state to a supercooled liquid state. On the other hand, regeneration air gains enthalpy in a sensible heat change process in the heat exchanger process.
When these mediums exchange heat to each other in the above described steps, the process can be approximated by a condensation heat transfer process at a constant temperature of 65° C. for the refrigerant, and by a sensible heat change process with an inlet temperature of 40° C. for the regeneration air, and this process can theoretically provide NTU (number of heat transfer unit) of about 1.7 and temperature effectiveness of 80% according to a characteristics of a crossflow type heat exchanger in which the refrigerant is mixed. Thus, outlet temperature of regeneration air is given by:
40+(65−40)×0.8=60° C.
so that regeneration air is heated to 60° C.
Therefore, as shown in
FIG. 9
, regeneration air stays in the heat exchanger
220
in a temperature range from 40° C. to 60° C., and the refrigerant exchanges heat with such regeneration air in a state of disorderly temperature distribution. Therefore, refrigerant liquid having its lowest enthalpy at the exit of the condenser cannot always contact with the regeneration air at the inlet at its lowest temperature of 40° C., and supposedly contacts with the regeneration air at the average regeneration air temperature of 50° C. Assuming that refrigerant is supercooled by the heat transferred from the area corresponding to 10% of the entire heat transfer area, which may be overestimated, NTU at this area is given by:
NTU=
1.7×0.1=0.17,
and since the temperature effectiveness is approximated by the formula:
&PHgr;=1−1
/exp
(
NTU
),
the temperature effectiveness is theoretically calculated as:
&PHgr;=1−1
/exp
(0.17)=0.156.
Thus, the temperature of the supercooled refrigerant liquid is:
65−(65−50)×0.156=62.7° C.
By actually calculating the ratio of enthalpy change for the supercooling effect by using the enthalpy value at 62.7° C., a value of 2.5% is obtained as shown in FIG.
9
. Therefore, heat transfer area was overestimated in the assumption for the above calculation, and the actual NTU is lower and the degree of supercooling is further lower so that the refrigerant liquid temperature will become a little higher than the above calculated value.
Thus, even when retaining the regeneration air of 40° C. as the lowest temperature for a cooling heat source, refrigerant liquid can be cooled only to 62.7° C. at the most, that is, the heat energy retained in the refrigerant cannot be transferred effectively to the air. Also, since the refrigerant at the inlet of the low-temperature heat source heat exchanger still has a high enthalpy, refrigerating effect in the low-temperature heat source heat exchanger becomes small. Accordingly, heating quantity of the cooling regeneration air and the cooling effect are smaller than the case where refrigerant liquid is cooled to 40° C., so that it is necessary to circulate a larger quantity of refrigerant for heating regeneration air so as to degrade the coefficient of performance. (As shown in
FIG. 8
, if heating quantity of the heat pump is taken as 100%, compressor power corresponding to heat quantity of 28% is necessary, and the refrigeration effect remains 72%.)
This invention has been made to solve the problems outlined above by providing an air conditioning system that can produce continual dehumidification of supply air and desiccant regeneration, by
Armstrong Westerman Hattori McLeland & Naughton LLP
Buiz Michael
Ebara Corporation
Jones Melvin
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