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Investigation on effect of indoor air distribution strategy on solar air-conditioning systems
Accepted Manuscript Investigation on effect of indoor air distribution strategy on solar air-conditioning systems
K.F. Fong, C.K. Lee, Z. Lin PII:
S0960-1481(18)30862-0
DOI:
10.1016/j.renene.2018.07.065
Reference:
RENE 10339
To appear in:
Renewable Energy
Received Date:
28 September 2017
Accepted Date:
14 July 2018
Please cite this article as: K.F. Fong, C.K. Lee, Z. Lin, Investigation on effect of indoor air distribution strategy on solar air-conditioning systems, Renewable Energy (2018), doi: 10.1016/j. renene.2018.07.065
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ACCEPTED MANUSCRIPT 1
Investigation on effect of indoor air distribution strategy on solar air-
2
conditioning systems
3 4
K.F. Fong*, C.K. Lee, Z. Lin
5 6
Division of Building Science and Technology, College of Science and Engineering, City
7
University of Hong Kong, Hong Kong, China
8
*Corresponding author.
9
Tel: 852-3442-8724, fax: 852-3442-0443
10
Email address: [email protected]
11
Post address: 83 Tat Chee Avenue, Kowloon Tong, Hong Kong
12 13
Abstract
14 15
Stratum ventilation (SV), a new indoor air distribution strategy, has been
16
promoted for applications in different building premises in recent years. Compared to the
17
conventional mixing ventilation (MV), the prominent advantage of SV is that indoor
18
thermal comfort can be satisfied with a relatively high supply air temperature, hence less
19
energy consumption in refrigeration. In solar air-conditioning, the energy performance
20
can also be facilitated by high-temperature cooling. As such, the potential of SV to be
21
involved in solar air-conditioning was evaluated. In this study, the solar air-conditioning
22
systems included solar absorption cooling system (SAbCS), solar adsorption cooling
23
system (SAdCS), solar desiccant cooling system (SDCS), hybrid solar absorption-
24
desiccant cooling system (HSAbDCS) and hybrid solar adsorption-desiccant cooling
25
system (HSAdDCS). Their performances using SV and MV were determined through
26
year-round dynamic simulation.
27
SAdCS, SDCS, HSAbDCS and HSAdDCS associated with SV could have 35%, 54%,
28
59%, 29% and 44% saving in the annual primary energy consumption for building in
29
subtropical climate respectively. Benchmarked with the conventional air-conditioning
30
system, they could have primary energy saving up to 30%. Consequently, solar air-
31
conditioning and SV can have synergetic merit in building application in hot and humid
32
city.
Compared to the counterpart using MV, SAbCS,
33 1
ACCEPTED MANUSCRIPT 34
Keywords: Solar air-conditioning; Stratum ventilation; Absorption cooling; Adsorption
35
cooling; Desiccant cooling; High-temperature cooling.
36 37
1.
Introduction
38 39
Stratum ventilation (SV), a new indoor air distribution strategy for small and
40
medium rooms, in which the supply terminals are installed at the side-walls at the
41
breathing zone of sedentary occupants, as shown in Fig. 1. SV was found effective to
42
achieve good thermal comfort [1-4], no matter with the uniformly or asymmetrically
43
distributed heat gains. SV could also have better indoor air quality for the occupants all
44
in sedentary position [5,6] or mixed with walking occupants [7,8]. SV can be coupled
45
with different conventional air side systems, like constant air volume system and fan coil
46
system. The thermal neutral temperature under stratum ventilation was also found able to
47
be raised to 27C [9], which is significantly higher than that under conventional MV at
48
the same supply air flow rate. As such, the year-round energy saving of conventional air-
49
conditioning system using SV was found to be at least 25% and 44% against those using
50
displacement ventilation and MV respectively [10] when applied to a sub-tropical region.
51
Actually SV could have further energy merits due to its pull-down performance [11] and
52
the optimized room air condition [12]. In life cycle assessment, SV was found to have
53
31.7% carbon emissions cut compared to the conventional MV [13].
54
55 56 57 58 59 60
Fig. 1. Contrast among supply air (SA) and return air (RA) arrangement of mixing ventilation (MV), displacement ventilation (DV) and stratum ventilation (SV) in a building zone. The supply air of MV, DV and SV are provided at the ceiling, floor and breathing zone levels respectively.
61
In the previous study [14], it was found that the solar adsorption cooling and the
62
solar desiccant cooling could not have any year-round energy merit when compared to the
63
conventional air-conditioning system under the context of MV. It might be helpful if the
2
ACCEPTED MANUSCRIPT 64
feature of high-temperature supply air in SV could help enhance the technical feasibility
65
of the solar adsorption cooling system. As such, it is necessary to review this again if the
66
air distribution strategy is changed from MV to SV. On the other hand, the energy saving
67
potential of solar air-conditioning systems could be remarkable when the strategy of high-
68
temperature cooling was applied based on the previous study [15,16]. Radiant cooling
69
was involved and higher supply temperature of the cooling medium could be used. In the
70
previous study, year-round energy saving of solar desiccant dehumidification system
71
serving stratum ventilated classrooms was assured [17]. As a result, the indoor air
72
distribution strategy SV can be considered for the solar air-conditioning systems, since its
73
required supply air temperature is higher than that used in MV. If sorption chillers are
74
used, the chilled water supply temperature can be higher and the coefficient of
75
performance (COP) can be enhanced. Therefore, the aim of this study is to conduct
76
performance evaluation of various solar air-conditioning systems using SV for building
77
application in the hot-humid climate. A comparative study will be conducted to appraise
78
different solar cooling systems using SV and MV, as well as the conventional air-
79
conditioning system using these two types of indoor air distribution strategies. In this
80
work, the solar cooling systems involved would cover a variety, including the solar
81
sorption cooling systems; the solar desiccant cooling system; and the hybrid solar
82
sorption-desiccant cooling systems.
83 84
2.
System Description
2.1
Solar sorption cooling systems
85 86 87 88
Solar sorption cooling systems include the solar absorption cooling system
89
(SAbCS) and the solar adsorption cooling system (SAdCS). Fig. 2 depicts the schematic
90
diagram of solar adsorption cooling system, in which adsorption chiller is the core of air-
91
conditioning.
92
instead. Solar-thermal collectors acquire the heat that is stored in the hot water (HW)
93
storage tank. Thermal energy is then used to drive the sorption chiller, with supplement
94
of auxiliary heater in need. Chilled water from chiller is delivered to the air handling
95
unit, and conditioned air serves the building zone through the strategy of stratum
96
ventilation. Supply air is provided at the level of breathing zone, while return air is
97
removed at the ceiling level. Outdoor air (OA) is furnished to the building zone and
In solar absorption cooling system, absorption chiller would be used
3
ACCEPTED MANUSCRIPT 98
exhaust air (EA) is dumped away. Cooling tower is involved for heat rejection of the
99
chiller. With respect to the composition of supply air, both SAbCS and SAdCS are
100
classified as return-air design.
101 Regenerative Water Pump
Solar-thermal Collectors HW HW Pump Storage Tank
Cooling Tower
EA
RA
C D
SA
Building Zone
Aux. Heater
-
OA
Air Handling Unit
E
Ad d
Cooling Water Pump
Adsorption Chiller
Chilled Water Pump
102 103 104 105 106
Fig. 2. Schematic diagram of solar adsorption cooling system for stratum ventilation. (C: condenser; D: desorber; Ad: adsorber; E: evaporator)
107
2.2
Solar desiccant cooling system
108 109
Fig. 3 shows the schematic diagram of solar desiccant cooling system (SDCS) for
110
stratum ventilation. The desiccant cooling unit mainly includes the desiccant wheel, the
111
rotary heat exchanger (RHE), the evaporative coolers in supply and exhaust air streams,
112
the supply air fan (SAF) and the exhaust air fan (EAF). Heat is drawn from the hot water
113
storage tank through the desiccant water pump (DWP) to the regenerative heating coil for
114
desiccant wheel, with supplement of auxiliary heater when necessary. The capacity of the
115
heating coil is governed by a three-way heating coil valve (HCV). Since the outdoor air
116
is fully treated to become supply air, an enhanced thermal comfort and indoor air quality
117
can thus be maintained at the breathing zone in the building zone. The SDCS is classified
118
as a full-outdoor-air design.
119
4
ACCEPTED MANUSCRIPT Desiccant Water Pump
Solar-thermal Collectors HW HW Pump Storage Tank
Aux. Heater
Heating Coil Valve
RA EAF Evap. Cooler
SA
Evap. Cooler
Building Zone SAF
120 121 122 123 124 125
EA +
OA
Rotary Desiccant Wheel Heat Exchanger
Fig. 3. Schematic diagram of solar desiccant cooling system using full outdoor air for stratum ventilation. 2.3
Hybrid solar sorption-desiccant cooling systems
126 127
With the load sharing approach, a hybrid solar sorption-desiccant cooling system
128
can be generated.
As such, hybrid solar absorption-desiccant cooling system
129
(HSAbDCS) or hybrid solar adsorption-desiccant cooling system (HSAdDCS) for stratum
130
ventilation system is developed. The sorption chiller is mainly responsible for the zone
131
sensible load, while the desiccant cooling unit for the zone latent load. In this way,
132
individual control of the zone temperature and humidity can be adopted. More important,
133
the problem of insufficient latent cooling due to the higher supply air temperature usually
134
found in SV can be solved. Fig. 4 depicts the configuration of HSAbDCS, in which
135
absorption chiller is employed. A common hot water storage tank is used to provide the
136
driving heat for both the sorption chiller and the desiccant cooling unit, but separate
137
auxiliary heaters are offered for supplement purpose. In this case, large portion of return
138
air can be directed to mix with the treated outdoor air before entering the cooling coil in
139
the desiccant cooling unit. Both HSAbDCS and HSAdDCS are also classified as return-
140
air design.
141
5
ACCEPTED MANUSCRIPT Regenerative Water Pump
Aux. Heater
Solar-thermal Collectors HW HW Pump Storage Tank DWP RA
Cooling Tower
Aux. Heater
HCV EAF
SA -
Building Zone
+
D W
OA
G
C
Absorption Chiller
Ab E
OAF
SAF
Desiccant Cooling Unit
142 143 144 145 146 147
R H E
EA
Chilled Water Pump
Cooling Water Pump
Fig. 4. Schematic diagram of hybrid solar absorption-desiccant cooling system for stratum ventilation. (G: generator; C: condenser; Ab: absorber; E: evaporator) 3.
Methodology of study and analysis
3.1
Year-round dynamic system simulation
148 149 150 151
In this study, year-round dynamic system simulation was conducted with
152
reference to the previous works for the solar sorption cooling systems and the solar
153
desiccant cooling system in [14]; and the hybrid solar sorption-desiccant cooling systems
154
in [15,16]. As such, the various dynamic system models are briefly introduced in the
155
following sub-sections. The component-based dynamic simulation platform TRNSYS
156
[18] was used in this study. TRNSYS offers a user-friendly interactive platform for
157
building a system in which each piece of equipment in the system was represented by a
158
component and linked up as desired. TRNSYS provide many built-in components like
159
the building zone, solar radiation, solar collectors, fans, pumps, coils, cooling towers,
160
controllers, etc. which are commonly used in an air-conditioning system and well
161
documented. Besides, it also allows users to develop their own component models. In
162
this study, new component models were created for the adsorption chiller, absorption
163
chiller, vapor compression chiller and desiccant wheel based on interpolation of
164
performance data generated from respective mathematical models as briefed in the sub-
165
sections that followed.
166
perform the multi-dimensional linear interpolation which can be used in the newly
167
developed component models.
168
combinations of the selected operating conditions, the performance of the component at
TRNSYS includes a built-in subroutine (DynamicData) to By building a performance data file for different
6
ACCEPTED MANUSCRIPT 169
any operating conditions within the range limits of each operating condition covered by
170
the data file could then be computed. By tracking the year-round weather data, the
171
dynamic behavior of the building zone and the air-conditioning system (including the air
172
and water in different circuits of the system) could be fully accounted for. This was
173
particularly useful and necessary in the investigation of the system performance during
174
the low-load period when intermittent operation of the cooling equipment was expected
175
although not always adopted in similar analysis found in the literature. In order to
176
evaluate the system performance for the building in hot-humid climate, the typical
177
weather data of the subtropical Hong Kong from Chan et al. [19] was involved. System
178
simulations were made for one year basis with a simulation time step of 6 minutes.
179 180
3.1.1
Building zone
181 182
In this study, the function of building zone was a typical office, which was 14 m
183
(L) 14 m (W) 3.6 m (H). In general, the building design was based on the local
184
design practice [20]. The U-values of external wall, window and flat roof were 2.73
185
Wm-2K-1, 5.68 Wm-2K-1 and 0.39 Wm-2K-1 respectively. The shading coefficient of
186
window was 0.25 and the window to wall ratio was 0.5.
187
contributed by the 24 sedentary occupants, the lighting load of 17 W/m2 and the
188
equipment load of 25 W/m2. The occupancy schedule was from 8:00 a.m. to 6:00 p.m.
189
The outdoor air demand was 10 Litre/s/person.
190
conditions of 32.8 C and 71% RH; and the design indoor conditions of 25.5 C and 60%
191
RH, the zone cooling load and ventilation load were 20 kW and 9 kW respectively. For
192
the stratum ventilation to be applied in the building zone, Lin et al. [10] postulated the
193
design methodology in the variation of temperature inside the building zone. For the
194
variation of the humidity ratio, it is assumed to follow the same profile as the
195
temperature.
The internal load was
According to the design outdoor
196 197
3.1.2
Sorption chillers and parasitic equipment
198 199
The absorption chiller was single-effect using the refrigerant and the absorbent of
200
water and lithium bromide respectively, and its simulation model was based on that of
201
Kim and Infante Fereira [21]. Constant overall heat transfer values were employed for
7
ACCEPTED MANUSCRIPT 202
the generator, absorber, condenser and evaporator. The simultaneous heat and mass
203
transfer at the generator and absorber was computed based on energy and mass balance.
204
Regarding the condenser and the evaporator, each coil was divided into two or three parts
205
depending on the states of the refrigerant. An iterative method was then employed to
206
calculate the coil performance.
207
maintained a constant degree of superheat of the refrigerant (steam) at the absorber inlet.
208
The overall performance of the absorption chiller was then determined based on energy
209
and refrigerant flow balance in an iterative manner. The adsorption chiller applied the
210
refrigerant and the adsorbent of water and silica gel respectively, its model was developed
211
by Cho & Kim [22]. A two-bed system was considered in which one bed worked as an
212
adsorber and the other as a desorber in an alternate manner. Constant overall heat transfer
213
values were assumed for the adsorber, desorber, condenser and evaporator.
214
effectiveness calculated based on fixed temperature on one side was adopted when
215
calculating the water temperature change across the four component chambers.
216
general, the auxiliary heater and the regenerative water pump for the sorption chiller were
217
controlled by a thermostat based on the chilled water return temperature. A part-load
218
controller was used to reset the auxiliary heater. The chiller and the cooling water pump
219
were called in operation when the auxiliary heater controller was switched on and that the
220
regenerative water temperature was higher than a preset minimum. The cooling tower
221
was additionally governed by a thermostat based on the cooling water temperature
222
leaving the chiller during operation. The chilled water pump would be running during the
223
entire daily operating period.
A thermostatic expansion valve was used which
An In
224 225
In the HSAbDCS or the HSAdDCS, the sorption chiller was used to handle the
226
zone sensible load and the auxiliary cooling capacity required to cool the hot
227
dehumidified outdoor air to the zone temperature.
228
temperature of SV allows a higher chilled water supply temperature to be used, the design
229
chilled water return temperature was 21 C while the design condenser water entering
230
temperature was 30 C. The flow rates for the chilled, cooling and regenerative water
231
pumps were selected based on the temperature difference of 5 C. The flow rate for the
232
hot water pump was the sum of those for the desiccant and regenerative water pumps.
233
The capacity of the auxiliary heater was sized according to the worst situation that all the
234
regenerative heat required by the chiller was provided by the heater.
8
Since the higher supply air
ACCEPTED MANUSCRIPT 235 236
3.1.3
Desiccant cooling unit
237 238
In the desiccant cooling unit, the one-dimensional desiccant wheel model of
239
Zhang et al. [23] was used.
The flow channel through the desiccant wheel was
240
discretized into numerous segments. The simultaneous heat and mass transfer across the
241
flow channel was computed segment-by-segment which followed the air flow direction
242
(depending on whether the flow channel was in the process or regenerative air side) in an
243
iterative manner. Silica gel was employed as the desiccant material in this study. The
244
flow rate of the desiccant water pump was designed according to the 10 C hot water
245
temperature difference of the heating coil in the desiccant cooling unit. For the desiccant
246
cooling unit in the solar desiccant cooling system, a thermostat was applied to control the
247
auxiliary heater based on the zone temperature. When the controller was “on”, the
248
desiccant water pump and the associated auxiliary heater would be called in during the
249
daily operating period. The remaining components, including the desiccant wheel, the
250
heating coil, the rotary heat exchanger and the evaporative coolers, would operate if the
251
desiccant water temperature was higher than the preset minimum temperature of 59 C.
252
The supply air fan and the exhaust air fan would function throughout the daily operating
253
period. The capacity of the auxiliary heater was sized according to the worst situation
254
that all the regenerative heat required by the desiccant wheel was provided by the heater.
255 256
For the desiccant cooling unit in the hybrid cooling system design, the control
257
approach is different. A proportional controller was used to determine the operation of
258
the desiccant cooling unit based on the zone humidity ratio. When the controller was
259
“on”, the desiccant water pump and the auxiliary heater would be called in. The rest
260
components, such as the desiccant wheel, the heating coil, the rotary heat exchanger and
261
the exhaust air fan, would function if the desiccant water temperature was higher than the
262
preset minimum temperature. The outdoor air fan and the supply air fan would function
263
throughout the daily operating period.
264
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ACCEPTED MANUSCRIPT 266
3.1.4
Solar energy collection
267 268
Evacuated tube collectors were used, and the model was developed based on the
269
efficiency of HRS [24]. The total net area of the solar collectors was 100 m2, which was
270
based on the assumption that only the roof (same size as the building zone or 196 m2)
271
could be used to install the solar collectors and that suitable space should be allowed
272
between rows of solar collectors to avoid shading. Meanwhile, the size of the hot water
273
storage tank was 5 m3. The hot water pump, which was used to circulate water between
274
the solar collectors and the storage tank, would be switched on when the temperature at
275
the solar collector outlet exceeded the water temperature inside the hot water tank by 5
276
C.
277 278
3.1.5
Conventional air-conditioning system
279 280
In the conventional air-conditioning system (CAS), vapor compression chiller was
281
used and the modeling approach outlined by Lee [25] was adopted. A constant polytropic
282
coefficient was adopted for the compressor. For the condenser, evaporator and expansion
283
valve, similar approaches as those used for the absorption chiller were employed. The
284
overall performance of the vapor compression chiller was determined based on energy
285
and refrigerant charge balance in an iterative manner. In this study, R134a was selected
286
as the refrigerant. The design parameters for the compression chiller were determined
287
based on the chilled water return temperature of 13 C with a rated COP. This would
288
allow the improvement of the chiller performance to be identified along with the increase
289
of the chilled water supply temperature.
290 291 292
Table 1 presents the design conditions for different systems under analysis, while Table 2 summarizes the major design parameters of the various air-conditioning systems.
293
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ACCEPTED MANUSCRIPT 295
Table 1. Design conditions for various solar air-conditioning system for SV. Design parameter
296 297 298 299
Supply air flow rate (kgs-1) Supply air temperature (C) Fresh air flow rate (kgs-1) Desiccant load (kW) Desiccant regeneration load/temperature (kW/C) Chiller load (kW) Driving load/temp. of absorption chiller (kW/C) COP of absorption chiller Driving load/temp. of adsorption chiller (kW/C) COP of adsorption chiller COP of vapor compression chiller Remark: NA refers to “not applicable”.
SAbCS / SAdCS / CASRA 2.15 19.9 0.288 NA NA 24.4 30.2/76 0.828 48.5/75 0.513 3.389
SDCS
HSAbDCS / HSAdDCS
CASOA
2.15 19.9 2.15 59.5 40.6/79 NA NA NA NA NA NA
2.15 19.9 0.288 7.0 6.4/81 19.8 24.4/79 0.823 39.8/77 0.501 NA
2.15 19.9 2.15 NA NA 52.7 NA NA NA NA 3.083
Table 2. Summary of design parameter values used in different air-conditioning systems for SV. Design parameter
SAbCS / SAdCS / CASRA
SDCS
Absorption chiller Regenerative water flow rate (kgs-1) Condenser water flow rate (kgs-1) Absorber water flow rate (kgs-1) Chilled water flow rate (kgs-1) Solution volume flow rate at absorber outlet (ls-1)
1.7 1.3 1.7 1.2 0.1
NA NA
Adsorption chiller Regenerative water flow rate (kgs-1) Condenser water flow rate (kgs-1) Adsorption chamber water flow rate (kgs-1) Vapor compression chiller Condenser water flow rate (kgs-1) Chilled water flow rate (kgs-1) Chilled water stream Chilled water pump power (kW)
HSAbDCS / HSAdDCS
CASOA
NA NA
NA NA
1.25 1.1 1.25 0.95 0.07
2.6 1.3 2.6
NA NA NA
1.8 1.1 1.9
NA NA NA
1.6 1.2
NA NA
NA NA
3.5 2.6
0.300
NA
0.215
0.716
NA
1.806 (Ab) 2.361 (Ad)
3.333
0.556 (Ab) 0.726 (Ad)
1.026
0.478 (Ab) 0.680 (Ad)
0.858
NA NA
Cooling water stream 2.361 (Ab) 3.056 (Ad) 1.250 (CAS) 0.726 (Ab) 0.940 (Ad) 0.385 (CAS) 0.680 (Ab) 0.793
Cooling tower air volume flow rate (m3s-1)
Cooling tower fan power (kW)
Cooling water pump power (kW)
11
ACCEPTED MANUSCRIPT (Ad) 0.341 (CAS) Desiccant wheel Outer diameter of desiccant wheel (m) Desiccant wheel power consumption (kW)
NA NA
1.6 0.2
0.6 0.1
NA NA
Rotary heat exchanger Temperature effectiveness power consumption (kW)
NA NA
0.8 0.2
0.8 0.1
NA NA
Supply air stream Supply air fan power (kW)
1.918
1.918
1.534
1.534
Outdoor air stream Fresh air fan power (kW)
NA
NA
0.185
NA
Exhaust air stream Exhaust air flow rate (kgs-1) Exhaust air fan power (kW)
NA NA
2.15 1.918
0.259 0.166
1.93 0.690
NA
Desiccant water stream Desiccant water pump flow rate (kgs-1) Desiccant water pump power (kW)
NA
NA NA
0.97 0.068
0.15 0.01
NA NA
0.207 (Ab) 0.316 (Ad)
NA
0.149 (Ab) 0.221 (Ad)
NA
0.97
1.4 (Ab) 1.95 (Ad)
NA
0.063
0.091 (Ab) 0.192 (Ad)
NA
Regenerative water stream Regenerative water pump power (kW)
Hot water stream Hot water flow rate (kgs-1) Hot water pump power (kW)
300 301
Remark: NA refers to “not applicable”.
302
3.2
1.7 (Ab) 2.6 (Ad) 0.139 (Ab) 0.217 (Ad)
Performance metrics
303 304
The primary performance metric of the solar air-conditioning systems is solar
305
fraction (SF), as defined in Eq. (1). SF measures the contribution of the regenerative heat
306
that comes from the solar thermal gain Qsolar. A higher SF is favorable, showing a lower
307
proportion of the regenerative heat from the auxiliary heater Qah.
308 309
SF
Qsolar Qsolar Qah
(1)
310
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ACCEPTED MANUSCRIPT 311
The second performance metric is COP, its definition depends on the type of air-
312
conditioning system. The COP of absorption chiller COPAb or adsorption chiller COPAd
313
is expressed in Eq. (2). For desiccant cooling, COPDC in Eq. (3) is for SDCS, while
314
COPhDC in Eq. (4) for the desiccant cooling part in the hybrid system like HSAbDCS or
315
HSAdDCS. The COP of the conventional vapor compression chiller COPVCC is shown in
316
Eq. (5).
317 318
COPAb or COPAd
Qc Qrh
(2)
319 320
COPDC
H sa H oa ,i Qhc
(3)
COPhDC
H oa , saci H oa ,i Qhc
(4)
COPVCC
Qc W
(5)
321 322 323 324 325 326
where Qc is cooling capacity, Qrh is regenerative heat to sorption chiller, Hoa,saci is
327
enthalpy of outdoor air at supply air coil inlet, Hoa,i is enthalpy of outdoor air intake, Qhc
328
is regenerative heat of heating coil, and W is electrical power input.
329 330
The third performance metric is primary energy consumption. There were two
331
forms of energy sources involved in the systems under study, one was electricity for
332
electrical equipment, while another was gas for auxiliary heating. In each type of air-
333
conditioning system, all kinds of energy consumptions including those of parasitic
334
equipment (like pumps, fans and cooling tower) would be covered. An energy efficiency
335
of 33% was assumed for the electric power generation and transmission/distribution in
336
relation to the primary energy input. The auxiliary heaters ran directly on primary energy
337
with a combustion efficiency of 90%.
338
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ACCEPTED MANUSCRIPT 340
4.
Results and discussions
4.1
Year-round performances
341 342 343 344
Table 3 consolidates the year-round performance of various air-conditioning
345
systems. For those solar air-conditioning systems with return-air design, there were 35%,
346
54%, 29% and 44% total primary energy saving in the SAbCSSV, the SAdCSSV, the
347
HSAbDCSSV and the HSAdDCSSV respectively, when compared with the corresponding
348
MV systems. A higher percentage reduction in the primary energy consumption could be
349
achieved by those cooling systems with an adsorption chiller. This was because COP of
350
adsorption chiller was lower when compared to those with an absorption chiller or a
351
vapor compression chiller. Hence, with the same reduction in the chiller capacity by
352
adopting a SV design, the decrease in the driving energy was more substantial for an
353
adsorption chiller. This could be reflected by the fact that the percentage rise in the SF
354
was also more significant for those solar air-conditioning systems with an adsorption
355
chiller. For the conventional air-conditioning systems, the CASRA,SV could have 38%
356
primary energy saving against the CASRA,MV, showing the same fact that SV had clear
357
energy merit on air-conditioning system.
358 359
Table 3 also summarizes the year-round performances of the various full-outdoor-
360
air systems with the two air distribution strategies. Clearly, the employment of SV
361
reduced the total primary energy consumption substantially, with 59% for the SDCSSV
362
and 54% for the CASOA,SV. The higher percentage saving for the SDCSSV was due to the
363
reduction of the supply air flow from 3.36 kgs-1 (according to [14]) to 2.15 kgs-1. The
364
resulting drop in the design system load was 39% while it was only 27% for the CASOA,SV.
365
Besides, the primary energy consumption from the fans was also lowered with the
366
reduction in the air flow. In other words, the employment of SV enhanced the benefit a
367
SDCS over a CAS in the full-outdoor air systems.
368 369
When compared with the conventional system using MV, no matter the CASRA,MV
370
or the CASOA,MV, all the solar air-conditioning systems using SV became technically
371
feasible with a maximum of 59% primary energy saving (in SDCSSV) and minimum 35%
372
(in HSAdDCSSV). If compared with the CASRA,SV or the CASOA,SV, the primary energy
14
ACCEPTED MANUSCRIPT 373
saving of the solar air-conditioning systems would be less, ranging from 6.4% (in
374
SAdCSSV) to 30% (in SAbCSSV), except that the HSAdDCSSV was technically infeasible
375
due to a slightly higher primary energy consumption. Actually the HSAdDCSSV needed
376
tangible energy in desiccant cooling in order to provide a more satisfactory indoor
377
relative humidity for the building zone, which will be discussed in Section 4.2.
378 379 380
Table 3. Summary of year-round performances of various air-conditioning systems with two indoor air distribution strategies. System
381 382 383
Average Tzone (C)
Average RHzone (%)
Total primary energy consumption (kW)
Average SF
Average COP
System with returnair design SAbCSSV SAbCSMV [14] SAdCSSV SAdCSMV [14]
0.98 0.82 0.89 0.58
25.6 24.5 25.5 24.5
70.4 58.6 72.0 58.9
32,218 49,425 43,184 93,692
HSAbDCSSV
0.93
25.5
57.3
34,191
HSAbDCSMV [26]
0.81
24.9
58.3
48,122
HSAdDCSSV
0.82
25.5
57.8
48,677
HSAdDCSMV [26]
0.59
0.787 (Ab) 0.763 (Ab) 0.485 (Ad) 0.437 (Ad) 0.822 (Ab) 0.750 (DC) 0.779 (Ab) 0.904 (DC) 0.509 (Ad) 0.812 (DC) 0.460 (Ad) 0.919 (DC) 3.646 (VCC) 3.195 (VCC)
24.9
58.4
87,215
25.2 24.8
73.8 58.9
46,126 74,431
1.008 (DC) 1.059 (DC) 3.363 (VCC) 3.177 (VCC)
24.6 23.6 24.1 24.5
72.7 66.7 77.8 66.0
52,775 128,052 59,274 130,053
CASRA,SV NA CASRA,MV [26] NA System with fulloutdoor-air design SDCSSV 0.90 SDCSMV [14] 0.55 CASOA,SV NA CASOA,MV [26] NA Remark: NA refers to “not applicable”.
4.2
Monthly performances
4.2.1
Systems with return-air design
384 385 386 387
The performances of the solar air-conditioning systems could be evaluated more
388
deeply through the annual profiles. Fig. 5 depicts those of primary energy consumption
389
of various air-conditionings with return-air design.
390
profiles of all solar air-conditioning systems were under that of CASRA,MV, and those
391
related to absorption cooling were under that of CASRA,SV. For the primary energy
392
profiles associated to adsorption cooling, it depended on the months concerned. It was
15
Apparently, the primary energy
ACCEPTED MANUSCRIPT 393
obvious that the HSAdDCSSV had more months (actually altogether 7 months) with
394
primary energy consumption higher than the CASRA,SV, resulting in technically infeasible
395
among the various solar air-conditioning designs under study.
396
although there were not primary energy saving in 5 months, it was feasible due to the
397
lower annual primary energy consumption. For most solar cooling systems investigated,
398
a local minimum existed in July for the total primary energy consumption. This could be
399
explained by the fact that the collected solar energy was high in July which out-weighed
400
the increased cooling demand and led to a decrease in total primary energy consumption.
For the SAdCSSV,
401
402 403 404 405
Fig. 5. Variation of monthly primary energy consumption for different air-conditioning systems with return-air design.
406
Besides the energy viewpoint, it was also important to evaluate the systems from
407
the indoor conditions acquired. Since it was found that some of the average zone relative
408
humidities were greater than 60% as shown in Table 5. In this regard, Figs. 6 and 7
409
compare the profiles of the monthly-averaged zone temperature and relative humidity of
410
different air-conditioning systems with return-air design. The adoption of SV in the non-
411
hybrid air-conditioning systems resulted in a high zone relative humidity during the peak-
412
load season. On the other hand, the hybrid solar sorption-desiccant cooling systems with
413
independent temperature and humidity control maintained the zone relative humidity well
414
within the satisfactory level. The non-hybrid SV systems offered a lower primary energy
415
consumption which would deem to have less satisfactory indoor environment during the
416
peak-load season.
417 16
ACCEPTED MANUSCRIPT
418 419 420 421
Fig. 6. Variation of monthly-averaged zone temperature for different air-conditioning systems with return-air design.
422 423 424 425
Fig. 7. Variation of monthly-averaged zone relative humidity for different airconditioning systems with return-air design.
426
4.2.2
Systems with full-outdoor-air design
427 428
Fig. 8 illustrates the monthly energy performance of the air-conditioning systems
429
with full-outdoor-air design. Obviously, the CASOA,MV had comparatively high primary
430
energy consumption around the year. Although the SDCSSV could have lower total
431
primary energy consumption than the CASOA,SV, it had higher energy consumption other
432
than the peak-load season. This was because the SDCSSV could also provide indoor
433
relative humidity closer to the design value even in that period of a year.
17
ACCEPTED MANUSCRIPT 434
435 436 437 438
Fig. 8. Variation of monthly primary energy consumption for different full-outdoor-air systems.
439
Figs. 9 and 10 show the profiles of the monthly-averaged temperature and relative
440
humidity for the various full-outdoor-air systems analyzed. Compared with the results for
441
CASOA, the zone relative humidity was higher with the employment of SV.
442
condition was unsatisfactory with CASOA,SV where the peak averaged relative humidity
443
could rise up to 87% during the peak-load period. The situation was improved with
444
SDCSSV where the peak averaged relative humidity only reached around 80%.
445
Nevertheless, the SDCSSV was still considered better than the conventional system for
446
full-outdoor-air application when the energy consumption was of primary concern.
447
18
The
ACCEPTED MANUSCRIPT
449 450 451 452
Fig. 9. Variation of monthly-averaged zone temperature for different full-outdoor-air systems.
453 454 455 456
Fig. 10. Variation of monthly-averaged zone relative humidity for different full-outdoorair systems.
457
So far the analysis made was based on the weather data of Hong Kong for office
458
buildings. The results (in particular the relative performances of different system designs)
459
could also be applied to other sub-tropical regions with a hot summer and cool winter like
460
the nearby density-populated cities in the Guangdong province of Mainland China for
461
buildings with similar loading intensity, occupant density and operating schedules.
462
19
ACCEPTED MANUSCRIPT 464
5.
Conclusion
465 466
For building application in the subtropical climate, it was found that the indoor air
467
distribution strategy had a definite effect on the performance of solar air-conditioning
468
systems. For those systems with return air design, the SAbCSSV, the SAdCSSV, the
469
HSAbDCSSV and the HSAdDCSSV had 35%, 54%, 29% and 44% saving respectively in
470
the annual primary energy consumption, as compared to the counterpart using MV. For
471
the full-outdoor-air design, the SDCSSV would have 59% primary energy saving. When
472
benchmarked with the CASMV, they had remarkable primary energy saving from 35% to
473
59%. The solar air-conditioning systems could generally have tangible primary energy
474
saving from 6.4% to 30%, except the HSAdDCSSV.
475 476
Under MV, the solar adsorption cooling system and the solar desiccant cooling
477
system could not have energy merit over the conventional provision as found in the
478
previous study. When SV was applied, these two systems could have lower year-round
479
primary energy consumption, which made them to be technically feasible options in solar
480
air-conditioning. Of course, the indoor conditions should be reviewed at the same time.
481
Although the solar sorption cooling systems using SV could generally have better energy
482
performances, they would cause the indoor RH around 80% for office application in the
483
peak-load season. For the building types with high latent cooling load, the hybrid system
484
with DCS would be more capable.
485 486
Acknowledgement
487 488
This work described in this paper is fully supported by a grant from the Research
489
Grants Council of the Hong Kong Special Administrative Region, China (Project No.
490
CityU 11200315).
491 492
Nomenclature
493 494
COP
Coefficient of performance
495
Hoa,saci
Enthalpy of outdoor air at supply air coil inlet (kW)
496
Hoa,i
Enthalpy of outdoor air intake (kW)
497
Hsa
Enthalpy of supply air (kW) 20
ACCEPTED MANUSCRIPT 498
Qah
Regenerative heat from auxiliary heating (kW)
499
Qc
Cooling capacity (kW)
500
Qhc
Regenerative heat of heating coil (kW)
501
Qrh
Regenerative heat to sorption chiller (kW)
502
Qsolar
Solar thermal gain (kW)
503
RHzone
Zone air relative humidity (%)
504
SF
Solar fraction
505
Tzone
Zone dry-bulb air temperature (C)
506 507
Subscript
508
Ab
Absorption chiller
509
Ad
Adsorption chiller
510
DC
Desiccant cooling
511
MV
Mixing ventilation
512
OA
Outdoor air
513
RA
Return air
514
SV
Stratum ventilation
515
VCC
Vapor compression chiller
516 517
Abbreviations
518 519
Ab
Absorber
520
Ad
Adsorber
521
C
Condenser
522
CAS
Conventional air-conditioning system
523
CWP
Cooling water pump
524
D
Desorber
525
DC
Desiccant cooling
526
DWP
Desiccant water pump
527
E
Evaporator
528
EA
Exhaust air
529
EAF
Exhaust air fan
530
OA
Outdoor air
21
ACCEPTED MANUSCRIPT 531
OAF
Outdoor air fan
532
HCV
Heating coil valve
533
HSAbDCS
Hybrid solar absorption-desiccant cooling system
534
HSAdDCS
Hybrid solar adsorption-desiccant cooling system
535
HW
Hot water
536
HWP
Hot water pump
537
MV
Mixing ventilation
538
RA
Return air
539
RWP
Regenerative water pump
540
SA
Supply air
541
SAbCS
Solar absorption cooling system
542
SAdCS
Solar adsorption cooling system
543
SAF
Supply air fan
544
SDCS
Solar desiccant cooling system
545
SV
Stratum ventilation
546
VCC
Vapor compression chiller
547 548
References
549 550
[1]
Lin, Z., Chow, T.T., Tsang, C.F., Fong, K.F. and Chan, L.S., Stratum ventilation -
551
a potential solution to elevated indoor temperature. Building and Environment,
552
44(11), 2009, pp. 2256-2269.
553
[2]
Cheng, Y., Fong, M.L., Yao, T., Lin, Z., Fong, K.F., Uniformity of stratum-
554
ventilated thermal environment and thermal sensation. Indoor Air, 24, 2014, pp.
555
521-532.
556
[3]
Cheng, Y., Lin, Z., Fong, A.M.L., Effects of temperature and supply airflow rate
557
on thermal comfort in a stratum-ventilated room. Building and Environment, 92,
558
2015, pp. 269-277.
559
[4]
Huan, C., Wang, F.H., Lin, Z., Wu, X.Z., Ma, Z.J., Wang, Z.H., Zhang, L.H., An
560
experimental investigation into stratum ventilation for the cooling of an office
561
with asymmetrically distributed heat gains. Building and Environment, 110, 2016,
562
pp. 76-88.
22
ACCEPTED MANUSCRIPT 563
[5]
Tian, L., Lin, Z., Wang, Q.W. and Liu, J., Comparison of gaseous contaminant
564
diffusion under stratum ventilation and under displacement ventilation. Building
565
and Environment, 45(9), 2010, pp. 2035-2046.
566
[6]
Lin, Z., Wang, J., Yao, T., Chow, T.T., Investigation into anti-airborne infection
567
performance of stratum ventilation. Building and Environment, 54, 2012, pp. 29-
568
38.
569
[7]
570 571
Wu, W., Lin, Z., An experimental study of the influence of a walking occupant on three air distribution methods. Building and Environment, 85, 2015, pp. 211-219.
[8]
Wu, W., Lin, Z., Experimental study of the influence of a moving manikin on
572
temperature profile and carbon dioxide distribution under three air distribution
573
methods. Building and Environment, 87, 2015, pp. 142-153.
574
[9]
Fong, M.L., Lin, Z., Fong, K.F., Chow, T.T. and Yao, T., Evaluation of thermal
575
comfort conditions in a classroom with three ventilation methods. Indoor Air,
576
21(3), 2011, pp. 231-239.
577
[10]
Lin, Z., Lee, C.K., Fong, S., Chow, T.T., Yao, T. and Chan, A.L.S., Comparison
578
of annual energy performances with different ventilation methods for cooling.
579
Energy and Buildings, 43(1), 2011, pp. 130-136.
580
[11]
Wang, X., Lin, Z., An experimental investigation into the pull-down performances
581
with different air distributions. Applied Thermal Engineering, 91, 2015, pp. 151-
582
162.
583
[12]
Zhang, S., Cheng, Y., Fang, Z., Huan, C., Lin, Z., Optimization of room air
584
temperature in stratum-ventilated rooms for both thermal comfort and energy
585
saving. Applied Energy, 204, 2017, pp. 420-431.
586
[13]
587 588
Fong, M.L., Lin, Z., Fong, K.F., Hanby, V., Greenough, R., Life cycle assessment for three ventilation methods. Building and Environment, 116, 2017, pp. 73-88.
[14]
Fong, K.F., Chow, T.T., Lee, C.K., Lin, Z. and Chan, L.S., Comparative study of
589
different solar cooling systems for buildings in subtropical city, Solar Energy,
590
84(2), 2010, pp. 227-244.
591
[15]
Fong, K.F., Lee, C.K., Chow, T.T., Lin, Z. and Chan, L.S., Solar hybrid air-
592
conditioning system for high temperature cooling in subtropical city. Renewable
593
Energy, 35(11), 2010, pp. 2439-2451.
594 595
[16]
Fong, K.F., Chow, T.T., Lee, C.K., Lin, Z., Chan, L.S., Solar hybrid cooling system for high-tech offices in subtropical climate – Radiant cooling by
23
ACCEPTED MANUSCRIPT 596
absorption refrigeration and desiccant dehumidification. Energy Conversion and
597
Management, 52(8-9), 2011, pp. 2883-2894.
598
[17]
Lee, C.K., Fong, K.F., Lin, Z., Chow, T.T., Year-round energy saving potential of
599
stratum ventilated classrooms with temperature and humidity control. HVAC&R
600
Research, 19, 2013, pp. 986-991.
601
[18]
602 603
TRNSYS 16, a TRaNsient SYstem Simulation program, the Solar Energy Laboratory, University of Wisconsin-Madison, 2006.
[19]
Chan, A.L.S., Chow, T.T., Fong, S.K.F. and Lin, J.Z., Generation of a typical
604
meteorological year for Hong Kong. Energy Conversion and Management, 47(1),
605
2006, pp. 87-96.
606
[20]
Guidelines on Performance-based Building Energy Code, EMSD, Hong Kong
607
SAR
608
http://www.emsd.gov.hk/filemanager/en/content_725/PBBEC_Guidelines_2007.p
609
df (last accessed on 25 June 2018).
610
[21]
Government.
Website:
Kim, D.S., and Infante Ferreira C.A., Analytical modeling of steady state single-
611
effect absorption cycles. International Journal of Refrigeration, 31(6), 2008, pp.
612
1012-1020.
613
[22]
614 615
Cho, S.H. and Kim, J.N., Modeling of a silica gel/water adsorption-cooling system. Energy 17(9), 1992, pp. 829-839.
[23]
Zhang, X.J., Dai, Y.J. and Wang, R.Z., A simulation study of heat and mass
616
transfer in a honeycombed rotary desiccant dehumidifier. Applied Thermal
617
Engineering, 23(8), 2003, pp. 989-1003.
618
[24]
Hochschul Rapperswil of Switzerland, Test Report No. 264, 1997.
619
[25]
Lee, C.K., Dynamic performance of ground-source heat pumps fitted with
620
frequency inverters for part-load control. Applied Energy, 87(11), 2010, pp. 3507-
621
3513.
622
[26]
Fong, K.F. Fong, Lee, C.K., Lin, Z., Chow, T.T., Chan, L.S., Application
623
potential of solar air-conditioning systems for displacement ventilation. Energy
624
and Buildings, 43, 2011, pp. 2068-2076.
625
24
ACCEPTED MANUSCRIPT Highlights > Solar air-conditioning (AC) facilitated by high-temperature cooling. > Stratum ventilation (SV) can also provide comfort with higher supply air temperature. > Hence solar AC systems using SV evaluated against those using mixing ventilation (MV). > Compared to the counterpart using MV, energy saving up to 59%. > Benchmarked with the conventional air-conditioning system, energy saving up to 30%.
K.F. Fong, C.K. Lee, Z. Lin PII:
S0960-1481(18)30862-0
DOI:
10.1016/j.renene.2018.07.065
Reference:
RENE 10339
To appear in:
Renewable Energy
Received Date:
28 September 2017
Accepted Date:
14 July 2018
Please cite this article as: K.F. Fong, C.K. Lee, Z. Lin, Investigation on effect of indoor air distribution strategy on solar air-conditioning systems, Renewable Energy (2018), doi: 10.1016/j. renene.2018.07.065
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ACCEPTED MANUSCRIPT 1
Investigation on effect of indoor air distribution strategy on solar air-
2
conditioning systems
3 4
K.F. Fong*, C.K. Lee, Z. Lin
5 6
Division of Building Science and Technology, College of Science and Engineering, City
7
University of Hong Kong, Hong Kong, China
8
*Corresponding author.
9
Tel: 852-3442-8724, fax: 852-3442-0443
10
Email address: [email protected]
11
Post address: 83 Tat Chee Avenue, Kowloon Tong, Hong Kong
12 13
Abstract
14 15
Stratum ventilation (SV), a new indoor air distribution strategy, has been
16
promoted for applications in different building premises in recent years. Compared to the
17
conventional mixing ventilation (MV), the prominent advantage of SV is that indoor
18
thermal comfort can be satisfied with a relatively high supply air temperature, hence less
19
energy consumption in refrigeration. In solar air-conditioning, the energy performance
20
can also be facilitated by high-temperature cooling. As such, the potential of SV to be
21
involved in solar air-conditioning was evaluated. In this study, the solar air-conditioning
22
systems included solar absorption cooling system (SAbCS), solar adsorption cooling
23
system (SAdCS), solar desiccant cooling system (SDCS), hybrid solar absorption-
24
desiccant cooling system (HSAbDCS) and hybrid solar adsorption-desiccant cooling
25
system (HSAdDCS). Their performances using SV and MV were determined through
26
year-round dynamic simulation.
27
SAdCS, SDCS, HSAbDCS and HSAdDCS associated with SV could have 35%, 54%,
28
59%, 29% and 44% saving in the annual primary energy consumption for building in
29
subtropical climate respectively. Benchmarked with the conventional air-conditioning
30
system, they could have primary energy saving up to 30%. Consequently, solar air-
31
conditioning and SV can have synergetic merit in building application in hot and humid
32
city.
Compared to the counterpart using MV, SAbCS,
33 1
ACCEPTED MANUSCRIPT 34
Keywords: Solar air-conditioning; Stratum ventilation; Absorption cooling; Adsorption
35
cooling; Desiccant cooling; High-temperature cooling.
36 37
1.
Introduction
38 39
Stratum ventilation (SV), a new indoor air distribution strategy for small and
40
medium rooms, in which the supply terminals are installed at the side-walls at the
41
breathing zone of sedentary occupants, as shown in Fig. 1. SV was found effective to
42
achieve good thermal comfort [1-4], no matter with the uniformly or asymmetrically
43
distributed heat gains. SV could also have better indoor air quality for the occupants all
44
in sedentary position [5,6] or mixed with walking occupants [7,8]. SV can be coupled
45
with different conventional air side systems, like constant air volume system and fan coil
46
system. The thermal neutral temperature under stratum ventilation was also found able to
47
be raised to 27C [9], which is significantly higher than that under conventional MV at
48
the same supply air flow rate. As such, the year-round energy saving of conventional air-
49
conditioning system using SV was found to be at least 25% and 44% against those using
50
displacement ventilation and MV respectively [10] when applied to a sub-tropical region.
51
Actually SV could have further energy merits due to its pull-down performance [11] and
52
the optimized room air condition [12]. In life cycle assessment, SV was found to have
53
31.7% carbon emissions cut compared to the conventional MV [13].
54
55 56 57 58 59 60
Fig. 1. Contrast among supply air (SA) and return air (RA) arrangement of mixing ventilation (MV), displacement ventilation (DV) and stratum ventilation (SV) in a building zone. The supply air of MV, DV and SV are provided at the ceiling, floor and breathing zone levels respectively.
61
In the previous study [14], it was found that the solar adsorption cooling and the
62
solar desiccant cooling could not have any year-round energy merit when compared to the
63
conventional air-conditioning system under the context of MV. It might be helpful if the
2
ACCEPTED MANUSCRIPT 64
feature of high-temperature supply air in SV could help enhance the technical feasibility
65
of the solar adsorption cooling system. As such, it is necessary to review this again if the
66
air distribution strategy is changed from MV to SV. On the other hand, the energy saving
67
potential of solar air-conditioning systems could be remarkable when the strategy of high-
68
temperature cooling was applied based on the previous study [15,16]. Radiant cooling
69
was involved and higher supply temperature of the cooling medium could be used. In the
70
previous study, year-round energy saving of solar desiccant dehumidification system
71
serving stratum ventilated classrooms was assured [17]. As a result, the indoor air
72
distribution strategy SV can be considered for the solar air-conditioning systems, since its
73
required supply air temperature is higher than that used in MV. If sorption chillers are
74
used, the chilled water supply temperature can be higher and the coefficient of
75
performance (COP) can be enhanced. Therefore, the aim of this study is to conduct
76
performance evaluation of various solar air-conditioning systems using SV for building
77
application in the hot-humid climate. A comparative study will be conducted to appraise
78
different solar cooling systems using SV and MV, as well as the conventional air-
79
conditioning system using these two types of indoor air distribution strategies. In this
80
work, the solar cooling systems involved would cover a variety, including the solar
81
sorption cooling systems; the solar desiccant cooling system; and the hybrid solar
82
sorption-desiccant cooling systems.
83 84
2.
System Description
2.1
Solar sorption cooling systems
85 86 87 88
Solar sorption cooling systems include the solar absorption cooling system
89
(SAbCS) and the solar adsorption cooling system (SAdCS). Fig. 2 depicts the schematic
90
diagram of solar adsorption cooling system, in which adsorption chiller is the core of air-
91
conditioning.
92
instead. Solar-thermal collectors acquire the heat that is stored in the hot water (HW)
93
storage tank. Thermal energy is then used to drive the sorption chiller, with supplement
94
of auxiliary heater in need. Chilled water from chiller is delivered to the air handling
95
unit, and conditioned air serves the building zone through the strategy of stratum
96
ventilation. Supply air is provided at the level of breathing zone, while return air is
97
removed at the ceiling level. Outdoor air (OA) is furnished to the building zone and
In solar absorption cooling system, absorption chiller would be used
3
ACCEPTED MANUSCRIPT 98
exhaust air (EA) is dumped away. Cooling tower is involved for heat rejection of the
99
chiller. With respect to the composition of supply air, both SAbCS and SAdCS are
100
classified as return-air design.
101 Regenerative Water Pump
Solar-thermal Collectors HW HW Pump Storage Tank
Cooling Tower
EA
RA
C D
SA
Building Zone
Aux. Heater
-
OA
Air Handling Unit
E
Ad d
Cooling Water Pump
Adsorption Chiller
Chilled Water Pump
102 103 104 105 106
Fig. 2. Schematic diagram of solar adsorption cooling system for stratum ventilation. (C: condenser; D: desorber; Ad: adsorber; E: evaporator)
107
2.2
Solar desiccant cooling system
108 109
Fig. 3 shows the schematic diagram of solar desiccant cooling system (SDCS) for
110
stratum ventilation. The desiccant cooling unit mainly includes the desiccant wheel, the
111
rotary heat exchanger (RHE), the evaporative coolers in supply and exhaust air streams,
112
the supply air fan (SAF) and the exhaust air fan (EAF). Heat is drawn from the hot water
113
storage tank through the desiccant water pump (DWP) to the regenerative heating coil for
114
desiccant wheel, with supplement of auxiliary heater when necessary. The capacity of the
115
heating coil is governed by a three-way heating coil valve (HCV). Since the outdoor air
116
is fully treated to become supply air, an enhanced thermal comfort and indoor air quality
117
can thus be maintained at the breathing zone in the building zone. The SDCS is classified
118
as a full-outdoor-air design.
119
4
ACCEPTED MANUSCRIPT Desiccant Water Pump
Solar-thermal Collectors HW HW Pump Storage Tank
Aux. Heater
Heating Coil Valve
RA EAF Evap. Cooler
SA
Evap. Cooler
Building Zone SAF
120 121 122 123 124 125
EA +
OA
Rotary Desiccant Wheel Heat Exchanger
Fig. 3. Schematic diagram of solar desiccant cooling system using full outdoor air for stratum ventilation. 2.3
Hybrid solar sorption-desiccant cooling systems
126 127
With the load sharing approach, a hybrid solar sorption-desiccant cooling system
128
can be generated.
As such, hybrid solar absorption-desiccant cooling system
129
(HSAbDCS) or hybrid solar adsorption-desiccant cooling system (HSAdDCS) for stratum
130
ventilation system is developed. The sorption chiller is mainly responsible for the zone
131
sensible load, while the desiccant cooling unit for the zone latent load. In this way,
132
individual control of the zone temperature and humidity can be adopted. More important,
133
the problem of insufficient latent cooling due to the higher supply air temperature usually
134
found in SV can be solved. Fig. 4 depicts the configuration of HSAbDCS, in which
135
absorption chiller is employed. A common hot water storage tank is used to provide the
136
driving heat for both the sorption chiller and the desiccant cooling unit, but separate
137
auxiliary heaters are offered for supplement purpose. In this case, large portion of return
138
air can be directed to mix with the treated outdoor air before entering the cooling coil in
139
the desiccant cooling unit. Both HSAbDCS and HSAdDCS are also classified as return-
140
air design.
141
5
ACCEPTED MANUSCRIPT Regenerative Water Pump
Aux. Heater
Solar-thermal Collectors HW HW Pump Storage Tank DWP RA
Cooling Tower
Aux. Heater
HCV EAF
SA -
Building Zone
+
D W
OA
G
C
Absorption Chiller
Ab E
OAF
SAF
Desiccant Cooling Unit
142 143 144 145 146 147
R H E
EA
Chilled Water Pump
Cooling Water Pump
Fig. 4. Schematic diagram of hybrid solar absorption-desiccant cooling system for stratum ventilation. (G: generator; C: condenser; Ab: absorber; E: evaporator) 3.
Methodology of study and analysis
3.1
Year-round dynamic system simulation
148 149 150 151
In this study, year-round dynamic system simulation was conducted with
152
reference to the previous works for the solar sorption cooling systems and the solar
153
desiccant cooling system in [14]; and the hybrid solar sorption-desiccant cooling systems
154
in [15,16]. As such, the various dynamic system models are briefly introduced in the
155
following sub-sections. The component-based dynamic simulation platform TRNSYS
156
[18] was used in this study. TRNSYS offers a user-friendly interactive platform for
157
building a system in which each piece of equipment in the system was represented by a
158
component and linked up as desired. TRNSYS provide many built-in components like
159
the building zone, solar radiation, solar collectors, fans, pumps, coils, cooling towers,
160
controllers, etc. which are commonly used in an air-conditioning system and well
161
documented. Besides, it also allows users to develop their own component models. In
162
this study, new component models were created for the adsorption chiller, absorption
163
chiller, vapor compression chiller and desiccant wheel based on interpolation of
164
performance data generated from respective mathematical models as briefed in the sub-
165
sections that followed.
166
perform the multi-dimensional linear interpolation which can be used in the newly
167
developed component models.
168
combinations of the selected operating conditions, the performance of the component at
TRNSYS includes a built-in subroutine (DynamicData) to By building a performance data file for different
6
ACCEPTED MANUSCRIPT 169
any operating conditions within the range limits of each operating condition covered by
170
the data file could then be computed. By tracking the year-round weather data, the
171
dynamic behavior of the building zone and the air-conditioning system (including the air
172
and water in different circuits of the system) could be fully accounted for. This was
173
particularly useful and necessary in the investigation of the system performance during
174
the low-load period when intermittent operation of the cooling equipment was expected
175
although not always adopted in similar analysis found in the literature. In order to
176
evaluate the system performance for the building in hot-humid climate, the typical
177
weather data of the subtropical Hong Kong from Chan et al. [19] was involved. System
178
simulations were made for one year basis with a simulation time step of 6 minutes.
179 180
3.1.1
Building zone
181 182
In this study, the function of building zone was a typical office, which was 14 m
183
(L) 14 m (W) 3.6 m (H). In general, the building design was based on the local
184
design practice [20]. The U-values of external wall, window and flat roof were 2.73
185
Wm-2K-1, 5.68 Wm-2K-1 and 0.39 Wm-2K-1 respectively. The shading coefficient of
186
window was 0.25 and the window to wall ratio was 0.5.
187
contributed by the 24 sedentary occupants, the lighting load of 17 W/m2 and the
188
equipment load of 25 W/m2. The occupancy schedule was from 8:00 a.m. to 6:00 p.m.
189
The outdoor air demand was 10 Litre/s/person.
190
conditions of 32.8 C and 71% RH; and the design indoor conditions of 25.5 C and 60%
191
RH, the zone cooling load and ventilation load were 20 kW and 9 kW respectively. For
192
the stratum ventilation to be applied in the building zone, Lin et al. [10] postulated the
193
design methodology in the variation of temperature inside the building zone. For the
194
variation of the humidity ratio, it is assumed to follow the same profile as the
195
temperature.
The internal load was
According to the design outdoor
196 197
3.1.2
Sorption chillers and parasitic equipment
198 199
The absorption chiller was single-effect using the refrigerant and the absorbent of
200
water and lithium bromide respectively, and its simulation model was based on that of
201
Kim and Infante Fereira [21]. Constant overall heat transfer values were employed for
7
ACCEPTED MANUSCRIPT 202
the generator, absorber, condenser and evaporator. The simultaneous heat and mass
203
transfer at the generator and absorber was computed based on energy and mass balance.
204
Regarding the condenser and the evaporator, each coil was divided into two or three parts
205
depending on the states of the refrigerant. An iterative method was then employed to
206
calculate the coil performance.
207
maintained a constant degree of superheat of the refrigerant (steam) at the absorber inlet.
208
The overall performance of the absorption chiller was then determined based on energy
209
and refrigerant flow balance in an iterative manner. The adsorption chiller applied the
210
refrigerant and the adsorbent of water and silica gel respectively, its model was developed
211
by Cho & Kim [22]. A two-bed system was considered in which one bed worked as an
212
adsorber and the other as a desorber in an alternate manner. Constant overall heat transfer
213
values were assumed for the adsorber, desorber, condenser and evaporator.
214
effectiveness calculated based on fixed temperature on one side was adopted when
215
calculating the water temperature change across the four component chambers.
216
general, the auxiliary heater and the regenerative water pump for the sorption chiller were
217
controlled by a thermostat based on the chilled water return temperature. A part-load
218
controller was used to reset the auxiliary heater. The chiller and the cooling water pump
219
were called in operation when the auxiliary heater controller was switched on and that the
220
regenerative water temperature was higher than a preset minimum. The cooling tower
221
was additionally governed by a thermostat based on the cooling water temperature
222
leaving the chiller during operation. The chilled water pump would be running during the
223
entire daily operating period.
A thermostatic expansion valve was used which
An In
224 225
In the HSAbDCS or the HSAdDCS, the sorption chiller was used to handle the
226
zone sensible load and the auxiliary cooling capacity required to cool the hot
227
dehumidified outdoor air to the zone temperature.
228
temperature of SV allows a higher chilled water supply temperature to be used, the design
229
chilled water return temperature was 21 C while the design condenser water entering
230
temperature was 30 C. The flow rates for the chilled, cooling and regenerative water
231
pumps were selected based on the temperature difference of 5 C. The flow rate for the
232
hot water pump was the sum of those for the desiccant and regenerative water pumps.
233
The capacity of the auxiliary heater was sized according to the worst situation that all the
234
regenerative heat required by the chiller was provided by the heater.
8
Since the higher supply air
ACCEPTED MANUSCRIPT 235 236
3.1.3
Desiccant cooling unit
237 238
In the desiccant cooling unit, the one-dimensional desiccant wheel model of
239
Zhang et al. [23] was used.
The flow channel through the desiccant wheel was
240
discretized into numerous segments. The simultaneous heat and mass transfer across the
241
flow channel was computed segment-by-segment which followed the air flow direction
242
(depending on whether the flow channel was in the process or regenerative air side) in an
243
iterative manner. Silica gel was employed as the desiccant material in this study. The
244
flow rate of the desiccant water pump was designed according to the 10 C hot water
245
temperature difference of the heating coil in the desiccant cooling unit. For the desiccant
246
cooling unit in the solar desiccant cooling system, a thermostat was applied to control the
247
auxiliary heater based on the zone temperature. When the controller was “on”, the
248
desiccant water pump and the associated auxiliary heater would be called in during the
249
daily operating period. The remaining components, including the desiccant wheel, the
250
heating coil, the rotary heat exchanger and the evaporative coolers, would operate if the
251
desiccant water temperature was higher than the preset minimum temperature of 59 C.
252
The supply air fan and the exhaust air fan would function throughout the daily operating
253
period. The capacity of the auxiliary heater was sized according to the worst situation
254
that all the regenerative heat required by the desiccant wheel was provided by the heater.
255 256
For the desiccant cooling unit in the hybrid cooling system design, the control
257
approach is different. A proportional controller was used to determine the operation of
258
the desiccant cooling unit based on the zone humidity ratio. When the controller was
259
“on”, the desiccant water pump and the auxiliary heater would be called in. The rest
260
components, such as the desiccant wheel, the heating coil, the rotary heat exchanger and
261
the exhaust air fan, would function if the desiccant water temperature was higher than the
262
preset minimum temperature. The outdoor air fan and the supply air fan would function
263
throughout the daily operating period.
264
9
ACCEPTED MANUSCRIPT 266
3.1.4
Solar energy collection
267 268
Evacuated tube collectors were used, and the model was developed based on the
269
efficiency of HRS [24]. The total net area of the solar collectors was 100 m2, which was
270
based on the assumption that only the roof (same size as the building zone or 196 m2)
271
could be used to install the solar collectors and that suitable space should be allowed
272
between rows of solar collectors to avoid shading. Meanwhile, the size of the hot water
273
storage tank was 5 m3. The hot water pump, which was used to circulate water between
274
the solar collectors and the storage tank, would be switched on when the temperature at
275
the solar collector outlet exceeded the water temperature inside the hot water tank by 5
276
C.
277 278
3.1.5
Conventional air-conditioning system
279 280
In the conventional air-conditioning system (CAS), vapor compression chiller was
281
used and the modeling approach outlined by Lee [25] was adopted. A constant polytropic
282
coefficient was adopted for the compressor. For the condenser, evaporator and expansion
283
valve, similar approaches as those used for the absorption chiller were employed. The
284
overall performance of the vapor compression chiller was determined based on energy
285
and refrigerant charge balance in an iterative manner. In this study, R134a was selected
286
as the refrigerant. The design parameters for the compression chiller were determined
287
based on the chilled water return temperature of 13 C with a rated COP. This would
288
allow the improvement of the chiller performance to be identified along with the increase
289
of the chilled water supply temperature.
290 291 292
Table 1 presents the design conditions for different systems under analysis, while Table 2 summarizes the major design parameters of the various air-conditioning systems.
293
10
ACCEPTED MANUSCRIPT 295
Table 1. Design conditions for various solar air-conditioning system for SV. Design parameter
296 297 298 299
Supply air flow rate (kgs-1) Supply air temperature (C) Fresh air flow rate (kgs-1) Desiccant load (kW) Desiccant regeneration load/temperature (kW/C) Chiller load (kW) Driving load/temp. of absorption chiller (kW/C) COP of absorption chiller Driving load/temp. of adsorption chiller (kW/C) COP of adsorption chiller COP of vapor compression chiller Remark: NA refers to “not applicable”.
SAbCS / SAdCS / CASRA 2.15 19.9 0.288 NA NA 24.4 30.2/76 0.828 48.5/75 0.513 3.389
SDCS
HSAbDCS / HSAdDCS
CASOA
2.15 19.9 2.15 59.5 40.6/79 NA NA NA NA NA NA
2.15 19.9 0.288 7.0 6.4/81 19.8 24.4/79 0.823 39.8/77 0.501 NA
2.15 19.9 2.15 NA NA 52.7 NA NA NA NA 3.083
Table 2. Summary of design parameter values used in different air-conditioning systems for SV. Design parameter
SAbCS / SAdCS / CASRA
SDCS
Absorption chiller Regenerative water flow rate (kgs-1) Condenser water flow rate (kgs-1) Absorber water flow rate (kgs-1) Chilled water flow rate (kgs-1) Solution volume flow rate at absorber outlet (ls-1)
1.7 1.3 1.7 1.2 0.1
NA NA
Adsorption chiller Regenerative water flow rate (kgs-1) Condenser water flow rate (kgs-1) Adsorption chamber water flow rate (kgs-1) Vapor compression chiller Condenser water flow rate (kgs-1) Chilled water flow rate (kgs-1) Chilled water stream Chilled water pump power (kW)
HSAbDCS / HSAdDCS
CASOA
NA NA
NA NA
1.25 1.1 1.25 0.95 0.07
2.6 1.3 2.6
NA NA NA
1.8 1.1 1.9
NA NA NA
1.6 1.2
NA NA
NA NA
3.5 2.6
0.300
NA
0.215
0.716
NA
1.806 (Ab) 2.361 (Ad)
3.333
0.556 (Ab) 0.726 (Ad)
1.026
0.478 (Ab) 0.680 (Ad)
0.858
NA NA
Cooling water stream 2.361 (Ab) 3.056 (Ad) 1.250 (CAS) 0.726 (Ab) 0.940 (Ad) 0.385 (CAS) 0.680 (Ab) 0.793
Cooling tower air volume flow rate (m3s-1)
Cooling tower fan power (kW)
Cooling water pump power (kW)
11
ACCEPTED MANUSCRIPT (Ad) 0.341 (CAS) Desiccant wheel Outer diameter of desiccant wheel (m) Desiccant wheel power consumption (kW)
NA NA
1.6 0.2
0.6 0.1
NA NA
Rotary heat exchanger Temperature effectiveness power consumption (kW)
NA NA
0.8 0.2
0.8 0.1
NA NA
Supply air stream Supply air fan power (kW)
1.918
1.918
1.534
1.534
Outdoor air stream Fresh air fan power (kW)
NA
NA
0.185
NA
Exhaust air stream Exhaust air flow rate (kgs-1) Exhaust air fan power (kW)
NA NA
2.15 1.918
0.259 0.166
1.93 0.690
NA
Desiccant water stream Desiccant water pump flow rate (kgs-1) Desiccant water pump power (kW)
NA
NA NA
0.97 0.068
0.15 0.01
NA NA
0.207 (Ab) 0.316 (Ad)
NA
0.149 (Ab) 0.221 (Ad)
NA
0.97
1.4 (Ab) 1.95 (Ad)
NA
0.063
0.091 (Ab) 0.192 (Ad)
NA
Regenerative water stream Regenerative water pump power (kW)
Hot water stream Hot water flow rate (kgs-1) Hot water pump power (kW)
300 301
Remark: NA refers to “not applicable”.
302
3.2
1.7 (Ab) 2.6 (Ad) 0.139 (Ab) 0.217 (Ad)
Performance metrics
303 304
The primary performance metric of the solar air-conditioning systems is solar
305
fraction (SF), as defined in Eq. (1). SF measures the contribution of the regenerative heat
306
that comes from the solar thermal gain Qsolar. A higher SF is favorable, showing a lower
307
proportion of the regenerative heat from the auxiliary heater Qah.
308 309
SF
Qsolar Qsolar Qah
(1)
310
12
ACCEPTED MANUSCRIPT 311
The second performance metric is COP, its definition depends on the type of air-
312
conditioning system. The COP of absorption chiller COPAb or adsorption chiller COPAd
313
is expressed in Eq. (2). For desiccant cooling, COPDC in Eq. (3) is for SDCS, while
314
COPhDC in Eq. (4) for the desiccant cooling part in the hybrid system like HSAbDCS or
315
HSAdDCS. The COP of the conventional vapor compression chiller COPVCC is shown in
316
Eq. (5).
317 318
COPAb or COPAd
Qc Qrh
(2)
319 320
COPDC
H sa H oa ,i Qhc
(3)
COPhDC
H oa , saci H oa ,i Qhc
(4)
COPVCC
Qc W
(5)
321 322 323 324 325 326
where Qc is cooling capacity, Qrh is regenerative heat to sorption chiller, Hoa,saci is
327
enthalpy of outdoor air at supply air coil inlet, Hoa,i is enthalpy of outdoor air intake, Qhc
328
is regenerative heat of heating coil, and W is electrical power input.
329 330
The third performance metric is primary energy consumption. There were two
331
forms of energy sources involved in the systems under study, one was electricity for
332
electrical equipment, while another was gas for auxiliary heating. In each type of air-
333
conditioning system, all kinds of energy consumptions including those of parasitic
334
equipment (like pumps, fans and cooling tower) would be covered. An energy efficiency
335
of 33% was assumed for the electric power generation and transmission/distribution in
336
relation to the primary energy input. The auxiliary heaters ran directly on primary energy
337
with a combustion efficiency of 90%.
338
13
ACCEPTED MANUSCRIPT 340
4.
Results and discussions
4.1
Year-round performances
341 342 343 344
Table 3 consolidates the year-round performance of various air-conditioning
345
systems. For those solar air-conditioning systems with return-air design, there were 35%,
346
54%, 29% and 44% total primary energy saving in the SAbCSSV, the SAdCSSV, the
347
HSAbDCSSV and the HSAdDCSSV respectively, when compared with the corresponding
348
MV systems. A higher percentage reduction in the primary energy consumption could be
349
achieved by those cooling systems with an adsorption chiller. This was because COP of
350
adsorption chiller was lower when compared to those with an absorption chiller or a
351
vapor compression chiller. Hence, with the same reduction in the chiller capacity by
352
adopting a SV design, the decrease in the driving energy was more substantial for an
353
adsorption chiller. This could be reflected by the fact that the percentage rise in the SF
354
was also more significant for those solar air-conditioning systems with an adsorption
355
chiller. For the conventional air-conditioning systems, the CASRA,SV could have 38%
356
primary energy saving against the CASRA,MV, showing the same fact that SV had clear
357
energy merit on air-conditioning system.
358 359
Table 3 also summarizes the year-round performances of the various full-outdoor-
360
air systems with the two air distribution strategies. Clearly, the employment of SV
361
reduced the total primary energy consumption substantially, with 59% for the SDCSSV
362
and 54% for the CASOA,SV. The higher percentage saving for the SDCSSV was due to the
363
reduction of the supply air flow from 3.36 kgs-1 (according to [14]) to 2.15 kgs-1. The
364
resulting drop in the design system load was 39% while it was only 27% for the CASOA,SV.
365
Besides, the primary energy consumption from the fans was also lowered with the
366
reduction in the air flow. In other words, the employment of SV enhanced the benefit a
367
SDCS over a CAS in the full-outdoor air systems.
368 369
When compared with the conventional system using MV, no matter the CASRA,MV
370
or the CASOA,MV, all the solar air-conditioning systems using SV became technically
371
feasible with a maximum of 59% primary energy saving (in SDCSSV) and minimum 35%
372
(in HSAdDCSSV). If compared with the CASRA,SV or the CASOA,SV, the primary energy
14
ACCEPTED MANUSCRIPT 373
saving of the solar air-conditioning systems would be less, ranging from 6.4% (in
374
SAdCSSV) to 30% (in SAbCSSV), except that the HSAdDCSSV was technically infeasible
375
due to a slightly higher primary energy consumption. Actually the HSAdDCSSV needed
376
tangible energy in desiccant cooling in order to provide a more satisfactory indoor
377
relative humidity for the building zone, which will be discussed in Section 4.2.
378 379 380
Table 3. Summary of year-round performances of various air-conditioning systems with two indoor air distribution strategies. System
381 382 383
Average Tzone (C)
Average RHzone (%)
Total primary energy consumption (kW)
Average SF
Average COP
System with returnair design SAbCSSV SAbCSMV [14] SAdCSSV SAdCSMV [14]
0.98 0.82 0.89 0.58
25.6 24.5 25.5 24.5
70.4 58.6 72.0 58.9
32,218 49,425 43,184 93,692
HSAbDCSSV
0.93
25.5
57.3
34,191
HSAbDCSMV [26]
0.81
24.9
58.3
48,122
HSAdDCSSV
0.82
25.5
57.8
48,677
HSAdDCSMV [26]
0.59
0.787 (Ab) 0.763 (Ab) 0.485 (Ad) 0.437 (Ad) 0.822 (Ab) 0.750 (DC) 0.779 (Ab) 0.904 (DC) 0.509 (Ad) 0.812 (DC) 0.460 (Ad) 0.919 (DC) 3.646 (VCC) 3.195 (VCC)
24.9
58.4
87,215
25.2 24.8
73.8 58.9
46,126 74,431
1.008 (DC) 1.059 (DC) 3.363 (VCC) 3.177 (VCC)
24.6 23.6 24.1 24.5
72.7 66.7 77.8 66.0
52,775 128,052 59,274 130,053
CASRA,SV NA CASRA,MV [26] NA System with fulloutdoor-air design SDCSSV 0.90 SDCSMV [14] 0.55 CASOA,SV NA CASOA,MV [26] NA Remark: NA refers to “not applicable”.
4.2
Monthly performances
4.2.1
Systems with return-air design
384 385 386 387
The performances of the solar air-conditioning systems could be evaluated more
388
deeply through the annual profiles. Fig. 5 depicts those of primary energy consumption
389
of various air-conditionings with return-air design.
390
profiles of all solar air-conditioning systems were under that of CASRA,MV, and those
391
related to absorption cooling were under that of CASRA,SV. For the primary energy
392
profiles associated to adsorption cooling, it depended on the months concerned. It was
15
Apparently, the primary energy
ACCEPTED MANUSCRIPT 393
obvious that the HSAdDCSSV had more months (actually altogether 7 months) with
394
primary energy consumption higher than the CASRA,SV, resulting in technically infeasible
395
among the various solar air-conditioning designs under study.
396
although there were not primary energy saving in 5 months, it was feasible due to the
397
lower annual primary energy consumption. For most solar cooling systems investigated,
398
a local minimum existed in July for the total primary energy consumption. This could be
399
explained by the fact that the collected solar energy was high in July which out-weighed
400
the increased cooling demand and led to a decrease in total primary energy consumption.
For the SAdCSSV,
401
402 403 404 405
Fig. 5. Variation of monthly primary energy consumption for different air-conditioning systems with return-air design.
406
Besides the energy viewpoint, it was also important to evaluate the systems from
407
the indoor conditions acquired. Since it was found that some of the average zone relative
408
humidities were greater than 60% as shown in Table 5. In this regard, Figs. 6 and 7
409
compare the profiles of the monthly-averaged zone temperature and relative humidity of
410
different air-conditioning systems with return-air design. The adoption of SV in the non-
411
hybrid air-conditioning systems resulted in a high zone relative humidity during the peak-
412
load season. On the other hand, the hybrid solar sorption-desiccant cooling systems with
413
independent temperature and humidity control maintained the zone relative humidity well
414
within the satisfactory level. The non-hybrid SV systems offered a lower primary energy
415
consumption which would deem to have less satisfactory indoor environment during the
416
peak-load season.
417 16
ACCEPTED MANUSCRIPT
418 419 420 421
Fig. 6. Variation of monthly-averaged zone temperature for different air-conditioning systems with return-air design.
422 423 424 425
Fig. 7. Variation of monthly-averaged zone relative humidity for different airconditioning systems with return-air design.
426
4.2.2
Systems with full-outdoor-air design
427 428
Fig. 8 illustrates the monthly energy performance of the air-conditioning systems
429
with full-outdoor-air design. Obviously, the CASOA,MV had comparatively high primary
430
energy consumption around the year. Although the SDCSSV could have lower total
431
primary energy consumption than the CASOA,SV, it had higher energy consumption other
432
than the peak-load season. This was because the SDCSSV could also provide indoor
433
relative humidity closer to the design value even in that period of a year.
17
ACCEPTED MANUSCRIPT 434
435 436 437 438
Fig. 8. Variation of monthly primary energy consumption for different full-outdoor-air systems.
439
Figs. 9 and 10 show the profiles of the monthly-averaged temperature and relative
440
humidity for the various full-outdoor-air systems analyzed. Compared with the results for
441
CASOA, the zone relative humidity was higher with the employment of SV.
442
condition was unsatisfactory with CASOA,SV where the peak averaged relative humidity
443
could rise up to 87% during the peak-load period. The situation was improved with
444
SDCSSV where the peak averaged relative humidity only reached around 80%.
445
Nevertheless, the SDCSSV was still considered better than the conventional system for
446
full-outdoor-air application when the energy consumption was of primary concern.
447
18
The
ACCEPTED MANUSCRIPT
449 450 451 452
Fig. 9. Variation of monthly-averaged zone temperature for different full-outdoor-air systems.
453 454 455 456
Fig. 10. Variation of monthly-averaged zone relative humidity for different full-outdoorair systems.
457
So far the analysis made was based on the weather data of Hong Kong for office
458
buildings. The results (in particular the relative performances of different system designs)
459
could also be applied to other sub-tropical regions with a hot summer and cool winter like
460
the nearby density-populated cities in the Guangdong province of Mainland China for
461
buildings with similar loading intensity, occupant density and operating schedules.
462
19
ACCEPTED MANUSCRIPT 464
5.
Conclusion
465 466
For building application in the subtropical climate, it was found that the indoor air
467
distribution strategy had a definite effect on the performance of solar air-conditioning
468
systems. For those systems with return air design, the SAbCSSV, the SAdCSSV, the
469
HSAbDCSSV and the HSAdDCSSV had 35%, 54%, 29% and 44% saving respectively in
470
the annual primary energy consumption, as compared to the counterpart using MV. For
471
the full-outdoor-air design, the SDCSSV would have 59% primary energy saving. When
472
benchmarked with the CASMV, they had remarkable primary energy saving from 35% to
473
59%. The solar air-conditioning systems could generally have tangible primary energy
474
saving from 6.4% to 30%, except the HSAdDCSSV.
475 476
Under MV, the solar adsorption cooling system and the solar desiccant cooling
477
system could not have energy merit over the conventional provision as found in the
478
previous study. When SV was applied, these two systems could have lower year-round
479
primary energy consumption, which made them to be technically feasible options in solar
480
air-conditioning. Of course, the indoor conditions should be reviewed at the same time.
481
Although the solar sorption cooling systems using SV could generally have better energy
482
performances, they would cause the indoor RH around 80% for office application in the
483
peak-load season. For the building types with high latent cooling load, the hybrid system
484
with DCS would be more capable.
485 486
Acknowledgement
487 488
This work described in this paper is fully supported by a grant from the Research
489
Grants Council of the Hong Kong Special Administrative Region, China (Project No.
490
CityU 11200315).
491 492
Nomenclature
493 494
COP
Coefficient of performance
495
Hoa,saci
Enthalpy of outdoor air at supply air coil inlet (kW)
496
Hoa,i
Enthalpy of outdoor air intake (kW)
497
Hsa
Enthalpy of supply air (kW) 20
ACCEPTED MANUSCRIPT 498
Qah
Regenerative heat from auxiliary heating (kW)
499
Qc
Cooling capacity (kW)
500
Qhc
Regenerative heat of heating coil (kW)
501
Qrh
Regenerative heat to sorption chiller (kW)
502
Qsolar
Solar thermal gain (kW)
503
RHzone
Zone air relative humidity (%)
504
SF
Solar fraction
505
Tzone
Zone dry-bulb air temperature (C)
506 507
Subscript
508
Ab
Absorption chiller
509
Ad
Adsorption chiller
510
DC
Desiccant cooling
511
MV
Mixing ventilation
512
OA
Outdoor air
513
RA
Return air
514
SV
Stratum ventilation
515
VCC
Vapor compression chiller
516 517
Abbreviations
518 519
Ab
Absorber
520
Ad
Adsorber
521
C
Condenser
522
CAS
Conventional air-conditioning system
523
CWP
Cooling water pump
524
D
Desorber
525
DC
Desiccant cooling
526
DWP
Desiccant water pump
527
E
Evaporator
528
EA
Exhaust air
529
EAF
Exhaust air fan
530
OA
Outdoor air
21
ACCEPTED MANUSCRIPT 531
OAF
Outdoor air fan
532
HCV
Heating coil valve
533
HSAbDCS
Hybrid solar absorption-desiccant cooling system
534
HSAdDCS
Hybrid solar adsorption-desiccant cooling system
535
HW
Hot water
536
HWP
Hot water pump
537
MV
Mixing ventilation
538
RA
Return air
539
RWP
Regenerative water pump
540
SA
Supply air
541
SAbCS
Solar absorption cooling system
542
SAdCS
Solar adsorption cooling system
543
SAF
Supply air fan
544
SDCS
Solar desiccant cooling system
545
SV
Stratum ventilation
546
VCC
Vapor compression chiller
547 548
References
549 550
[1]
Lin, Z., Chow, T.T., Tsang, C.F., Fong, K.F. and Chan, L.S., Stratum ventilation -
551
a potential solution to elevated indoor temperature. Building and Environment,
552
44(11), 2009, pp. 2256-2269.
553
[2]
Cheng, Y., Fong, M.L., Yao, T., Lin, Z., Fong, K.F., Uniformity of stratum-
554
ventilated thermal environment and thermal sensation. Indoor Air, 24, 2014, pp.
555
521-532.
556
[3]
Cheng, Y., Lin, Z., Fong, A.M.L., Effects of temperature and supply airflow rate
557
on thermal comfort in a stratum-ventilated room. Building and Environment, 92,
558
2015, pp. 269-277.
559
[4]
Huan, C., Wang, F.H., Lin, Z., Wu, X.Z., Ma, Z.J., Wang, Z.H., Zhang, L.H., An
560
experimental investigation into stratum ventilation for the cooling of an office
561
with asymmetrically distributed heat gains. Building and Environment, 110, 2016,
562
pp. 76-88.
22
ACCEPTED MANUSCRIPT 563
[5]
Tian, L., Lin, Z., Wang, Q.W. and Liu, J., Comparison of gaseous contaminant
564
diffusion under stratum ventilation and under displacement ventilation. Building
565
and Environment, 45(9), 2010, pp. 2035-2046.
566
[6]
Lin, Z., Wang, J., Yao, T., Chow, T.T., Investigation into anti-airborne infection
567
performance of stratum ventilation. Building and Environment, 54, 2012, pp. 29-
568
38.
569
[7]
570 571
Wu, W., Lin, Z., An experimental study of the influence of a walking occupant on three air distribution methods. Building and Environment, 85, 2015, pp. 211-219.
[8]
Wu, W., Lin, Z., Experimental study of the influence of a moving manikin on
572
temperature profile and carbon dioxide distribution under three air distribution
573
methods. Building and Environment, 87, 2015, pp. 142-153.
574
[9]
Fong, M.L., Lin, Z., Fong, K.F., Chow, T.T. and Yao, T., Evaluation of thermal
575
comfort conditions in a classroom with three ventilation methods. Indoor Air,
576
21(3), 2011, pp. 231-239.
577
[10]
Lin, Z., Lee, C.K., Fong, S., Chow, T.T., Yao, T. and Chan, A.L.S., Comparison
578
of annual energy performances with different ventilation methods for cooling.
579
Energy and Buildings, 43(1), 2011, pp. 130-136.
580
[11]
Wang, X., Lin, Z., An experimental investigation into the pull-down performances
581
with different air distributions. Applied Thermal Engineering, 91, 2015, pp. 151-
582
162.
583
[12]
Zhang, S., Cheng, Y., Fang, Z., Huan, C., Lin, Z., Optimization of room air
584
temperature in stratum-ventilated rooms for both thermal comfort and energy
585
saving. Applied Energy, 204, 2017, pp. 420-431.
586
[13]
587 588
Fong, M.L., Lin, Z., Fong, K.F., Hanby, V., Greenough, R., Life cycle assessment for three ventilation methods. Building and Environment, 116, 2017, pp. 73-88.
[14]
Fong, K.F., Chow, T.T., Lee, C.K., Lin, Z. and Chan, L.S., Comparative study of
589
different solar cooling systems for buildings in subtropical city, Solar Energy,
590
84(2), 2010, pp. 227-244.
591
[15]
Fong, K.F., Lee, C.K., Chow, T.T., Lin, Z. and Chan, L.S., Solar hybrid air-
592
conditioning system for high temperature cooling in subtropical city. Renewable
593
Energy, 35(11), 2010, pp. 2439-2451.
594 595
[16]
Fong, K.F., Chow, T.T., Lee, C.K., Lin, Z., Chan, L.S., Solar hybrid cooling system for high-tech offices in subtropical climate – Radiant cooling by
23
ACCEPTED MANUSCRIPT 596
absorption refrigeration and desiccant dehumidification. Energy Conversion and
597
Management, 52(8-9), 2011, pp. 2883-2894.
598
[17]
Lee, C.K., Fong, K.F., Lin, Z., Chow, T.T., Year-round energy saving potential of
599
stratum ventilated classrooms with temperature and humidity control. HVAC&R
600
Research, 19, 2013, pp. 986-991.
601
[18]
602 603
TRNSYS 16, a TRaNsient SYstem Simulation program, the Solar Energy Laboratory, University of Wisconsin-Madison, 2006.
[19]
Chan, A.L.S., Chow, T.T., Fong, S.K.F. and Lin, J.Z., Generation of a typical
604
meteorological year for Hong Kong. Energy Conversion and Management, 47(1),
605
2006, pp. 87-96.
606
[20]
Guidelines on Performance-based Building Energy Code, EMSD, Hong Kong
607
SAR
608
http://www.emsd.gov.hk/filemanager/en/content_725/PBBEC_Guidelines_2007.p
609
df (last accessed on 25 June 2018).
610
[21]
Government.
Website:
Kim, D.S., and Infante Ferreira C.A., Analytical modeling of steady state single-
611
effect absorption cycles. International Journal of Refrigeration, 31(6), 2008, pp.
612
1012-1020.
613
[22]
614 615
Cho, S.H. and Kim, J.N., Modeling of a silica gel/water adsorption-cooling system. Energy 17(9), 1992, pp. 829-839.
[23]
Zhang, X.J., Dai, Y.J. and Wang, R.Z., A simulation study of heat and mass
616
transfer in a honeycombed rotary desiccant dehumidifier. Applied Thermal
617
Engineering, 23(8), 2003, pp. 989-1003.
618
[24]
Hochschul Rapperswil of Switzerland, Test Report No. 264, 1997.
619
[25]
Lee, C.K., Dynamic performance of ground-source heat pumps fitted with
620
frequency inverters for part-load control. Applied Energy, 87(11), 2010, pp. 3507-
621
3513.
622
[26]
Fong, K.F. Fong, Lee, C.K., Lin, Z., Chow, T.T., Chan, L.S., Application
623
potential of solar air-conditioning systems for displacement ventilation. Energy
624
and Buildings, 43, 2011, pp. 2068-2076.
625
24
ACCEPTED MANUSCRIPT Highlights > Solar air-conditioning (AC) facilitated by high-temperature cooling. > Stratum ventilation (SV) can also provide comfort with higher supply air temperature. > Hence solar AC systems using SV evaluated against those using mixing ventilation (MV). > Compared to the counterpart using MV, energy saving up to 59%. > Benchmarked with the conventional air-conditioning system, energy saving up to 30%.