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Experimental investigation on the performance of a solar powered lithium bromide–water absorption cooling system
Accepted Manuscript Title: Experimental investigation on the performance of a solar powered lithium bromide-water absorption cooling system Author: Ming Li, Chengmu Xu, Reda Hassanien Emam Hassanien, Yongfeng Xu, Binwei Zhuang PII: DOI: Reference:
S0140-7007(16)30234-1 http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.07.023 JIJR 3396
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
7-1-2015 11-7-2016 30-7-2016
Please cite this article as: Ming Li, Chengmu Xu, Reda Hassanien Emam Hassanien, Yongfeng Xu, Binwei Zhuang, Experimental investigation on the performance of a solar powered lithium bromide-water absorption cooling system, International Journal of Refrigeration (2016), http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.07.023. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Experimental investigation on the performance of a solar powered lithium bromide-water absorption cooling system
Ming Li1,1, Chengmu Xu1, Reda Hassanien Emam Hassanien1,2, Yongfeng Xu1,3, Binwei Zhuang1. 1
Solar Energy Research Institute, Yunnan Normal University, Kunming 650500, China.
2
Agricultural Engineering Department, Faculty of Agriculture, Cairo University, Cairo 12613, Egypt.
3
Zhejiang Solar Energy Product Quality Inspection Center, Haining, Zhejiang 314416, China.
Highlights
The cooling and space heating performance of the system was experimental investigated.
The cooling performance of the system influent by running temperature was analyzed.
The cooling performance of the system influent by cooling temperature was analyzed.
The improvement methods for the system were analyzed and discussed.
Abstract: The Performance of solar cooling absorption system needs further research, due to its poor coefficient of performance (COP).Therefore; this study investigated the performance of a 23 kW solar powered single-effect lithium bromide-water (LiBr-H2O) absorption cooling system. Furthermore, the space heating mode was also investigated and the improvement methods were analyzed and discussed. The cooling system was driven by a parabolic trough collector of 56 m2 aperture area and used for cooling a 102 m2 meeting room. Research results revealed that the chiller’s maximum instantaneous refrigeration coefficient (chiller efficiency) could be up to 0.6. Most of the time, in sunny and clear sky days the daily solar heat fraction ranged from 0.33 to 0.41 and the collectors field efficiency ranged from 0.35 to 0.45. At the same time, chiller efficiency was varied from 0.25 to 0.7 and the daily cooling COP was varied from 0.11 to 0.27, respectively. Keywords: Solar cooling; Single-effect absorption chiller; Lithium Bromide-water; Parabolic trough solar collector (PTC); Cooling performance
Nomenclature Symbols
1
Corresponding author. Tel./fax: +86 871 65517266.
E-mail address: [email protected] (M. Li).
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Ac
collection area of PTC, m2
cp
specific heat, J kg-1 K-1
COPs,av
Average COP of the whole system.
Ei
input solar radiation into PTC, W
Ei,tot
total input solar radiation energy into PTC, MJ
Ib
solar beam radiation, W m-2
mF
mass flow rate of water which flow passing by PTC array, kg s-1
mF,g
mass flow rate of water flow passing by the chiller’s generator, kg
mF,c
mass flow rates chilled water, kg s-1
mF,e
mass flow rates of cooling water, kg s-1
mw,t
water mass in the tank, kg
Pte,ins
instantaneous thermal power of PTC, W
Pte,ins,s
instantaneous power of heat collection system, W
Pr
the gained refrigeration quantity from the absorption chiller system, W
Pg
consumption thermal power of generator , W
Pe
absorption power from chilled water of water evaporate, W
Pc
output power into cooling water by absorber and condenser, W
Qtc,ins
the received useful solar radiation energy, W
Qp,loss
pipeline heat losses, W
Qt,loss
heat losses of the hot storage tank, W
Qr
total refrigeration quantity, MJ
Qh
the gained heat of the system, MJ
T1 – T15
the temperatures of probes in Fig. 1, ºC
Ta
corridor temperature (T15 in Fig. 1), ºC
Tr
indoor (meeting room) temperatures, ºC
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Greek symbols ΔT
temperature difference, ºC
ηte,ins
the instantaneous thermal efficiency
ηte,ins,s
the instantaneous thermal efficiency of heat collection system
ηr
thermodynamics coefficient
ηoel
the optical end loss ratio of PTC
ηr,av
the average refrigeration efficiency
θ
incidence angle, °
Abbreviations COP
coefficient of performance
LiBr-H2O
lithium bromide-water
PTC
parabolic trough solar collector(s)
1. Introduction The global climate changes resulted in an increasing of energy demand for traditional air conditioning and heating systems in buildings which, consumes more fossil fuel meanwhile, increasing the carbon emissions. Therefore, using solar energy for air conditioning becomes one of the promising approaches to reduce energy consumptions and negative environmental impacts from buildings. There are two main types of solar cooling namely solar-thermal-driven and solar-PV-driven air conditioning technologies. Solar-PV-driven system uses a conventional vapor compression air conditioning cycle in which the electrical input is provided by solar PV panels. The solar-thermal-driven utilizes solar thermal energy to power the generator of an absorption refrigeration systems and it has three cooling technologies such as, solar absorption refrigeration, solar adsorption refrigeration and solar ejector refrigeration[1, 2]. However, solar absorption refrigeration is the most matured technologies and it has been studied more extensively than other systems. The main advantage of the solar absorption cooling technology is that the coefficient of performance (COP) is higher than that of other thermally operated cycles [3]. Furthermore, in a comparison between the different solar electric, solar thermal from the point of view of energy efficiency and economic feasibility, solar electric and thermo-mechanical systems appear to be more expensive than thermal sorption systems. Absorption and adsorption are comparable in terms that adsorption chillers are more expensive and bulkier than absorption chillers. The total cost of a single-effect LiBr-H2O absorption system is estimated to be the lowest [4]. So due to relatively simple configuration and low requirements for heat sources, a numerous studies of solar absorption refrigeration system are using single-effect LiBr-H2O absorption chillers, recently [5-10]. Furthermore, Double-effect LiBr-H2O absorption chiller systems coefficients of performance are almost twice higher than those obtained with single-effect systems. with Double-effect systems it is possible to obtain
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coefficients of performance as high as 1.12 at condenser temperatures of 30 °C but they need generator temperatures higher than 140 °C to reach evaporator temperatures as low as -5 °C [1115]. But there are some high requirements for heat sources maybe more than 120°C. So Double-effect LiBr-H2O system is sophisticated construct and high investment and operation cost. Nevertheless, Avanessian and Ameri reported that the CO2 emission of the single-effect system was respectively about 1.9 and 1.7 times higher than direct-fired and hot-water double-effect [16]. The efficiency enhancement of the solar absorption cooling systems components is essential to increase the COP of the whole system [17, 18]. The performance evaluation of a 35 kW LiBr-H2O absorption machine driven by 72 m2 heat pipe evacuated tube collector with a gas backup system showed that the actual average COP of the system was 0.33 while the maximum and minimum values were 0.50 and 0.17 respectively. These results were obtained based on the average collector efficiency and the solar fraction of 0.55 and at an ambient temperature of 32°C[19]. Lu et al.[20] investigated one two-phase thermo-syphon silica gel-water solar adsorption chiller and LiBr-H2O absorption chiller with new medium Compound Parabolic Concentrator (CPC) solar collectors. Results revealed that the efficiency of the medium temperature evacuated-tube CPC solar collector can reach 0.5 when the hot water temperature is 125°C. The average solar COPs of absorption system was 0.19 [20]. On the other hand, Ali et al. [21] investigated the Performance assessment of an integrated free cooling and solar powered single-effect LiBr-H2O absorption chiller in Oberhausen (Germany). The plant includes 35.17 kW cooling absorption chiller, 108 m2 evacuated tube collectors, a 6.8 m3 hot water tank, a 1.5 m3 cold water tank and a 134 kW cooling tower. The results illustrated that the chiller efficiency in some cooling months could be up to 70% while it was about 25% during the 5 years period of the plant operation; the monthly average efficiency value of collectors varies from 34.1% up to 41.8% and the five-year average value is about 28.3%. Hang et al. [22] investigated the energy performance analysis of solar absorption cooling system, a 23 kW double-effect absorption chiller driven by a 54 m2 External Compound Parabolic Concentrator (XCPC) solar collectors. The results showed that the daily average collector efficiency changed between 36% and 39%. The average coefficient of performance (COP) of the LiBr absorption chiller was between 0.91 and 1.02 with an average value of 1.0, and the daily solar COP was approximately at 0.374. On other study, the performance analysis of a mini-type solar-powered absorption cooling system with a cooling capacity of 8 kW, solar collector’s area of 96 m2 and a water storage tank of 3 m3 was conducted. When LiBr-H2O used as the working pairs of the chiller, the experimental results revealed that the average values of predicted mean vote and predicted percentage of dissatisfied of the test room were 0.22 and 5.89, respectively, and the power consumption was reduced by 43.5%. Meanwhile, the theoretical model predicted that the solar radiation intensity has a great impact on the performance of the solar powered absorption cooling system compared with the ambient temperature[23]. Lazzarin et al. [24]reported their field test results for a solar cooling plant and gave the performance of a typical day in terms of energy collected and temperature of the solar collector. They indicated that both energy collected and temperature of the solar collector were significantly various during the day. Due to the solar input constantly varies. Subsequently, the performance of the system may deteriorate and the solar powered refrigeration system will not be able to work consistently[25]. Venegas et al.[26] investigated the influence of operational variables on the performance of a solar absorption cooling system. Results showed that the most important variables which effected on the daily cooling energy produced and the daily averaged solar COP are the amount of
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the collected solar energy, the wind velocity and wind direction. On the other hand, it was reported that the optimum mass flow rates of hot water passing through the generator on a double-effect absorption chiller with 100 t of cooling capacity and evacuated tube collector have an important role on reducing the auxiliary energy[12] .Subsequently, three alterative designs (heat storage, cold storage, and refrigerant storage) for 24-hour-operating solar-powered LiBr-H2O absorption air-conditioning systems were analyzed. The results indicate that continuously operating solar-powered LiBr-H2O absorption with refrigerant storage was the most suitable alternative design for a 24-hour cooling effect[27]. It was demonstrated that, to obtain the high coefficient of performance for a small size absorption chiller which was developed from an old out-of-order commercial chiller, this chiller should be operated at 85°C of hot water temperature supplied to the generator[28]. Therefore, Calise et al. [29]simulated different solar cooling systems to find out the appropriate operation parameters for maximizing the COP of solar cooling systems. However, there are many challenges are remaining for further studied to enhance the COP of the absorption refrigeration technology. This study investigates the experimental performance of a single-effect lithium bromide absorption refrigeration system driven by PTC for air conditioning. Furthermore, it discusses and analyzes appropriate methods for improving the cooling performance.
2. System description This experiment has been conducted in Kunming, China. The collector was a parabolic trough solar collector (PTC). Fig. 1 shows the schematic diagram for the main parts of the refrigeration experimental system which includes PTC array, hot water storage tank (with a supplemental water tank), single-effect lithium bromide absorption chiller (TX-23), cooling tower and blower coils installed in the meeting room. Fig. 1 also illustrates the positions of temperature test points. Table 1 shows the sensors of temperatures and its instruction. Fig. 2 shows the main parts of the refrigeration experimental system. The Main measurements and instruments accuracy of the monitoring system are shown in table 2. The parameters and heat efficiency values at different temperature of PTC array were shown in table 3. The heat receiver, glass-metal evacuated absorber tube, was the core component of PTC and the specifications were shown in table 2 too. The PTC array is made up with two PTCs in series and the array space between two PTCs was 5.6 m. Table 4 shows some parameters of refrigeration chiller. The capacity of hot water storage tank was 1 m3. Table 5 shows the cooling tower parameters, the same type of cooling water has been used in the cooling tower. The area of end air-conditioning room was 102 m2 and 3 Brower coils installed on the top of the meeting room (relevant parameters are shown in Table 6). The room is used to conference room which can accommodate more than 20 persons. The high of the room is 3.5m. And the length and the width are 13.6 m and 7.5 m, respectively. There are two windows with 10m2 transparent fiberglass in the north wall. The south wall nearby the corridor with two glass doors and the area of each door is 3 m2. And then, the last two walls are made up with reinforced concrete with 10cm thickness.
3. Energy conversion Analysis
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Fig. 3 shows the heat flow diagram of the absorption chiller, its main energy conversion processes are as follows: (1) The input solar radiation into PTC: E i I b Ac cos
(1) Where, Ib, Ac and θ are the solar beam radiation, the collection area and the incident angle of sun ray, respectively. (2) The useful energy transformed from the received solar radiation Qtc,ins. According to Fig. 1, Q tc , ins m F c p [(T2 T1 ) (T4 T3 )]
(2)
Where, mF is the mass flow rate of water flowing in the receiver of PTC array, and cp is the specific heat capacity at constant pressure. (3) The pipeline heat loss between PTC and hot water storage tank Qp,loss,1: Q p , loss ,1 m F c p [(T11 T1 ) (T2 T3 ) (T4 T12 )]
(3)
(4) The calculation of the heat loss of hot water storage tank Qt,loss is relatively complicated. The The heat loss of hot water tank was tested in accordance with ISO 9459 and EN12976. Before the experiment, the water in the tank was heated by PTC and the temperature was not low than 60°C. In order to ensure the temperature consistency in the tank and avoid tank thermal stratification, an external water loop was adopted. The water was pumped by hot water pump out of the tank in the bottom and back into the tank in the top and then the hot water was circulated through the external water pump. The pump and the circulation pipes were wrapped with insulation material, the thickness of which was no less than 3cm. There temperature sensors was arranged in equal intervals in the vertical direction inside of the hot water tank. When the values of there sensors were close, there are no thermal stratification in the tank. And then the circulation pump must be shut down. The hot water temperature was 63.6°C. The water was placed for 15 hours without any interfere. And then, the circulation pump was turn on once again and the hot water was circulated until the temperatures of sensors were the same nearly. The temperature was 49.8°C. In real experimental test we can use the simple formula to calculate it as follows: Pt, loss , av
m w ,t c p T t
(4)
Where, mw,t is the water mass in the tank, Δt is the continue time of experiment, and ΔT is the temperature different of water with no thermal stratification in the tank within the time of Δt. When, Δt→0, it can be gained the instantaneous heat loss power.
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(5) The pipeline heat loss between hot water storage tank and chiller: Q p , loss , 2 2 m F , g c p (T11 T5 )
(5)
Where, mF,g is the mass flow rate of water flow passing by the chiller’s generator. (6) The heat absorbed from chilled water in the evaporator of refrigeration chiller, Qp,loss,2 (this issue will be analyzed in the next section). (7) The output heat from absorber and condenser to cooling water, this part of heat can be equivalent to the output heat Qc from cooling tower. (8) Qp,a,3 is the heat absorbed of pipeline between chiller and meeting room which comes from the out surface of pipeline. (9) Qa,room is the heat absorbed of blower coils which comes from the air in the meeting room.
4. Thermal efficiency of heat collection system 4.1. Thermal efficiency of PTC array The instantaneous thermal power and instantaneous thermal efficiency are respectively as follows: Pte , ins m F c p [(T 2 T1 ) (T 4 T3 )] te , ins
Pte , ins I b Ac cos
(6) (7)
Because of the roof orientation factor the PTC array could shade by adjacent buildings after 4:30. Therefore, the experiments could only be conducted before 4:30. The experiments have been also conducted under the windy weather conditions, the tested peak wind velocity reached to 8 m s -1, which was largely influence to the thermal efficiency of PTC array under such weather condition. Because the focal line was frequently move out of the absorber tube under such high wind velocity. Fig. 4 gives the test results of thermal efficiency of PTC array. Obviously, when the irradiance changed from 0.80 to 0.90 kW m-2 the instantaneous thermal efficiency ranged from 0.50 to 0.65. at the same time, the system instantaneous thermal power was changed between 24 and 27 kW. In the calculations, the mass flow rate of water mF was tested as 0.602 kg s-1. 4.2. Energy losses from the pipelines Here, the test of energy losses of the pipeline between PTC and hot water storage tank, energy losses of the pipeline between hot water storage tank and chiller, and energy losses of the pipeline between chiller and meeting room were conducted.
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The heat conduction coefficient of thermal insulation layer of pipeline was between 0.037-0.040 W mK-1 and the out surface of the thermal insulation was wrapped with aluminum alloy layer. And the three types of pipelines heat loss are shown in Table 7. It is worth pointing out that because of the temperature of chilled water inside the tube was lower than that of the outside, the heat was absorbed from outside to inside of tubes. Therefore, the energy losses of the pipeline between chiller and meeting room were cooling losses, rather than heat losses. 1
4.3. Heat losses of hot water storage tank The heat loss of water storage tank was test. At first, the tank was filled with 688kg hot water. And 15 hours later, the temperature of water has been decreased from 63.6°C to 49.8 °C. It can be estimated that the average heat losses power was about 640 W. Obviously, the average power of heat losses increased with the increase of hot water temperature. 4.4. Thermal efficiency of heat collection system When the pipeline heat losses and heat losses of hot water tank are considered, the instantaneous power of heat collection system can be expressed as: Ptc , ins , s Ptc ,ins Pp ,loss Pt,loss
(8)
While the instantaneous thermal efficiency of heat collection system can be expressed as: te , ins , s
Ptc , ins Pp , loss Pt,loss I b Ac cos
(9)
Where, Pp,,loss and Pt,loss are pipelines heat losses and hot water tank heat losses, respectively. Thus when the solar beam radiation changed between 0.80 and 0.90 kW m-2, the instantaneous thermal efficiency and the instantaneous power ranged from 0.32 to 0.42 and from 17 to 21 kW, respectively. (Fig. 5).
5. Performance of the refrigeration system This experiment was conducted mainly to test and investigate the performance of the refrigeration system. During the experiment many parameters such as solar beam radiation, wind velocity, ambient temperature, inlet and outlet temperature of PTC, hot water storage tank, heating, chilled water, and cooling water were tested. 5.1. Energy analysis of refrigeration process for cooling chiller The thermodynamics coefficient (ηr) was considered as the economy evaluation which defined as r Pe Pg
(10)
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Where, Pe is the absorption power from chilled water of water evaporates, and Pg is the consumption of heat energy. According to Fig. 1, the consumption thermal power of generator is as follows: Pg m F , g c p (T 6 T5 )
(11)
The absorption power from chilled water of water evaporates is: Pe m F , e c p (T 7 T8 )
(12)
The output power into cooling water by absorber and condenser is as the following: Pc m F , c c p (T10 T9 )
(13)
In the three equations above, mF,g, mF,c and mF,e are mass flow rates of hot water, chilled water and cooling water, respectively. T5 - T10 are the inlet and outlet temperatures of hot water, chilled water and cooling water (see Fig.1 and Table 1). According to the analysis above, if ignored the power consumed by circulating water pump, the cycle refrigeration coefficient of the chiller (chiller efficiency or COP of the chiller) can be express as： r
Pe Pg
m F , e c p ( T 7 T8 ) m F , g c p (T 6 T5 )
m F , e ( T 7 T8 ) m F , g (T 6 T5 )
(14)
5.2. Experimental results The cooling performance of the absorption chiller for ten days of experiment was conducted. The experiment was conducted from April to May, 2014 at Yunnan Yi Tong Solar Science and Technology Co. Ltd (in Kunming, China). At first the water in tank must be heated by PTC. When the temperature of hot water in the tank reached more than 65°C, the refrigeration chiller was turned on by manual. In order to understand the chiller performance under different operating temperatures, the different water temperature is adopted to driven chiller and the refrigeration performance can be got and recorded. But in order to protect refrigeration chiller, the test experiment must be shut down when the hot water temperature in the tank descended to around 40 °C. All of the components are controlled by manual. During experiment, the test value of mF, mF,g and mF,c are 0.602 kg s-1, 1.36 kg s-1 and 0.90 kg s-1 respectively. The experimental results are shown table 8 and Figs. 6, 7, 8 and 9 are shown the changing curves of system performance parameters along with time. Where, the figures show the experimental results in two days and Table 8 gives the experimental results for refrigeration performance during 14 days.
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According to Fig. 6, during the experiment, the wind speed is high in most sunny and cloud days and the maximum wind velocity vw reached to 8 m s-1 (generally it ranged between 2 and 5 m s-1), which had a huge impact on the thermal efficiency of PTC system. Fig. 7 shows that when the operating temperature of refrigeration chiller (the temperature of hot water) increased from 39 ºC to 80ºC, the lowest temperature of chilled water could reach to 9 ºC. At the same time, the temperature of cooling water changed between 20 ºC and 35 ºC. The cooling process was disconnected. At the beginning of the experiment, the temperatures of hot water and chilled water decreased rapidly. After chiller operating steadily, the system performance parameters changed stably along with time. The higher the inlet temperature of hot water, the lower the outlet temperature of the chilled water is. Fig. 8 shows the changing curves of refrigeration coefficient and refrigeration power of chiller along with time. It shows that the refrigeration coefficient ηr increased along with the refrigeration quantity Pr increase of chiller. Fig. 9 shows that in the first thirty minutes after refrigeration chiller operating, the indoor temperature of meeting room decreased sharply. After reaching a certain value, the indoor temperature began to gradually stabilize. Sometimes, the temperature fluctuated slightly. The indoor temperature increased along with the ambient temperature. Table 8 shows the average value of performance coefficient of the refrigeration system during each experimental day. Where, Ei,tot is the total incident solar radiation energy (the total input solar radiation energy into PTC), Qr is the total refrigeration quantity, ηr,av is the average refrigeration efficiency and COPs,av is the average COP of the refrigeration system. The total incident solar radiation energy Ei,tot is given as E i , tot
t s , st
t s ,bt
Ac I b cos (1 oel )dt S
(15)
Where, ts,bt and ts,st are the time of experiment starting and finishing, respectively, ηoel is the optical end loss ratio of PTC (Xu et al., 2014). The total refrigeration quantity is given as Qr
t r , st
t r ,bt
m F , e c p (T7 T8 )dt
(16)
Where, tr,bt and tr,st are the time of refrigeration chiller operating and stopping, respectively. The average refrigeration efficiency expressed as r , av
Qr
t r , st t r ,bt
m F , g c p (T5 T6 )dt
(17)
COPs,av is shown as
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C O Ps , av
Qr
(18)
E i , tot
Table 8 illustrates that the average refrigeration efficiency (ηr,av) and the average COP (COPs,av) of the whole system were relatively low under cloudy weather conditions. In such weather condition, the solar radiation was relatively small. So the hot water could not be heated to high temperature. In addition, when the sun covered by cloud, the pump was running with fixed frequency all the time which resulted in more pipeline heat losses. Therefore, the average refrigeration efficiency and the average COP of the system were relatively low. In contrast, under clear weather conditions, the refrigeration efficiency and system COP were high. The experiment results revealed that the average refrigeration efficiency of chiller changed between 0.17 to 0.60, and the refrigeration coefficient of the whole system ranged from 0.11 to 0.27. 5.3. Space heating performance The solar refrigeration system located at the tropical areas allows the refrigeration applications to be undertaken all the year around. However, in high latitudes with four obvious seasons of the year, we need to maximize the utilization of the system by conducting the cooling mode in summer and heating mode in winter. Thereby the solar energy utilization efficiency will be improved. In energy conversion process, the system heating mode was relatively simple compared to the cooling mode and the heat gained by PTC can be expressed with The energy conversion process of the system was relatively simple in the heating mode compared to the cooling mode, and the gained heat of the system can be expressed as Eq. (8). In the heating mode, the refrigeration unit (cooling chiller) acted as the heat exchanger. The output heat power from the cooling chiller can be expressed as Eq. (12). The system heating performance are tested with experiment and the results are shown in table 9 and the system characteristic parameter variation curves are given in Fig. 10 and Fig. 11. Where, Fig. 10 gives the variations of the running temperature of the absorption chiller, Where T7 and T8 are the inlet and outlet temperatures of the heat exchanger, respectively and the measurement positions of T5 to T8 are shown in Fig. 1. Fig. 11 gives the changes of indoor temperatures. Where, Tr1, Tr2 and Tr3 are the indoor temperatures and Ta (T15 in Fig. 1) is the corridor temperature. It shows that the indoor temperatures increased very quickly at the first 10 minutes of heat supply process. Table 9 gives the heat collected by PTC under different weather conditions. It shows that the average gained heat rate was about 0.4. In the practical applications, the total gained heat of the system can be estimated as Q h te , ins , s E i , tot (19) 5.4. Improvement analysis and discussion
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The experimental results revealed that the operating temperature and cooling temperature had a huge influence on the cooling performance of cooling system. Fig. 12 shows the changes of refrigeration coefficient at different running temperatures. The refrigeration quantity was increasing along with the increase of the hot water temperature. At the same time, the refrigeration coefficient was also substantially increasing with the increase of the hot water temperature. Where, T5 is the hot water temperature at inlet as shown in Fig. 1 and Table 1. In absorption refrigeration system, the function of cooling water is to exclude the absorption heat in the absorber and condensation heat in the condenser. Therefore, the intensity degree of absorption and condensation is related to temperature difference between inlet and outlet of cooling water in condenser and absorber. Fig. 13 shows the cooling performance influent by cooling water at different water temperatures. Where, T9 and T10 are the inlet and outlet temperatures of cooling water, respectively (see Fig. 1 and Table 1). Because the cooling pump was working in intermittent, the temperature of cooling water was displays in the shape of fluctuating with time. It can be also seen that when the cooling water temperatures reached to the maximum value, the temperature differences between inlet and outlet of cooling water was about 1.4ºC, while when the cooling water temperature down to the minimum value, the temperature differences between inlet and outlet of cooling water was about 3ºC. Therefore, when the inlet temperature of cooling water was 21ºC, the chiller refrigeration coefficient was about 0.68, when the inlet temperature of cooling water was 27.6ºC the refrigeration coefficient dropped to about 0.39. The refrigeration coefficient of the system is very close to that of the reference [21]. And the operation model was similar to that mentioned in reference [28]. In other words, in order to increase the refrigeration efficiency and refrigeration quantity, the temperature of cooling water should be as low as possible. In order to increase the comprehensive efficiency of the whole system (including refrigeration efficiency and heating efficiency), a certain operation conditions should be conducted as followings. Firstly, the area of PTC should be increased. Subsequently, temperature of the hot water and the quantity of heat storage will be improved. Thus, the cooling chiller can be started in a short period of time and operated at a capacity of 23 kW for a long running time. Secondly, the energy losses of each part of energy conversion materials should be reduced as far as possible, such as pipeline heat loss, hot water tank heat loss and so on. According to Table 6 the heat losses of system pipeline were relatively high; the main reason is that the length of pipelines among PTC, hot water tank and cooling chiller was too large. Therefore, to reduce the pipelines heat losses, the length of these pipelines should be reduced as possible. In order to tackle the heat losses problem, every part of the heat transfer pipelines should be shortened. According to the experimental measurements, the length of pipeline between PTC and hot water tank can be reduced to 50 m; the pipeline between hot water tank and cooling chiller can be shortened to 5 m. Thus, the heat losses between PTC and hot water tank will be approximately decreased from 3.90 - 6.2 kW to 1.40 - 2.30 kW. In addition, the heat losses between hot water tank and the cooling chiller will be also reduced from 0.90 - 1.40 kW to 0.25 - 0.40 kW. Subsequently, the gained heat of the whole system will be increased by 3.15 - 4.90 kW and the average COP of the whole system will be more than 0.3. Because the system was located in a windy area, so in order to improve system performance, system design optimization
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work must be carried out. One of the ways is that both of hot water storage tank and the cooling chiller should be installed in a small room so as to minimize the convection heat loss of them. Thirdly, to increase the chiller refrigeration quantity Qr and the COP of the whole cycle, the temperature of cooling water should be as low as possible. So the way of continuous cooling should be used. According to Fig. 13, when continuous cooling is adopted, the refrigeration efficiency immediately raised to 0.7. On the contrary, in intermittent cooling model, the average refrigeration efficiency was only about 0.55. 6. Conclusion The cooling performance of the single-effect LiBr–H2O absorption chiller driven by PTC has been investigated. The results revealed that the chiller’s average refrigeration coefficient ηr,av was between 0.18 and 0.60, and the average COP of the whole refrigeration cycle COPs,av was from 0.11 to 0.27 under different weather conditions. Moreover, the refrigeration quantity of the refrigeration chiller has been increased with the increasing of heat temperature. Meanwhile the refrigeration coefficient also substantially increased with the rising of hot water temperature. The cooling water temperature could largely influence on the performance of the refrigeration chiller. The values of ηte,ins,s and Pte,ins,s were relatively low, due to the pipeline was too long and its heat loss was too high. In addition, the values of ηr,av and COPs,av were also relatively low because the hot water temperature was too low. Therefore, in order to improve the cooling performance of the existing system the length of heat transfer pipelines should be shorten to reduce the heat losses. Furthermore, area of PTC and the operating temperature of the aperture should be increased and the water temperature for cooling system should be as low as possible. Acknowledgement The present study was supported by National Natural Science Foundation, China (Grant No.: U1137605). References [1] Otanicar T, Taylor RA, Phelan PE. Prospects for solar cooling – An economic and environmental assessment. Solar Energy. 2012;86:1287-99. [2] Siddiqui MU, Said SAM. A review of solar powered absorption systems. Renewable and Sustainable Energy Reviews. 2015;42:93-115. [3] Hassan HZ, Mohamad AA. A review on solar cold production through absorption technology. Renewable and Sustainable Energy Reviews. 2012;16:5331-48. [4] Kim DS, Infante Ferreira CA. Solar refrigeration options- a state-of-the-art review. International Journal of Refrigeration. 2008;31:3-15. [5] Iranmanesh A, Mehrabian MA. Dynamic simulation of a single-effect LiBr-H2O absorption refrigeration cycle considering the effects of thermal masses. Energy and Buildings. 2013;60:47-59. [6] Izquierdo M, González-Gil A, Palacios E. Solar-powered single-and double-effect directly air-cooled LiBr-H2O absorption prototype built as a single unit. Applied Energy. 2014;130:7-19. [7] Lizarte R, Izquierdo M, Marcos JD, Palacios E. Experimental comparison of two solar-driven air-cooled LiBr-H2O absorption chillers: Indirect versus direct air-cooled system. Energy and Buildings. 2013;62:323-34. [8] Agyenim F, Knight I, Rhodes M. Design and experimental testing of the performance of an outdoor LiBr-H2O solar thermal absorption cooling system with a cold store. Solar Energy. 2010;84:735-44. [9] Lamine CM, Said Z. Energy Analysis of Single Effect Absorption Chiller (LiBr-H2O) in an Industrial Manufacturing of Detergent. Energy Procedia. 2014;50:105-12.
Page 13 of 24
[10] González-Gil A, Izquierdo M, Marcos JD, Palacios E. Experimental evaluation of a direct air-cooled lithium bromide–water absorption prototype for solar air conditioning. Applied Thermal Engineering. 2011;31:3358-68. [11] Li Z, Ye X, Liu J. Performance analysis of solar air cooled double effect LiBr-H2O absorption cooling system in subtropical city. Energy Conversion and Management. 2014;85:302-12. [12] Iranmanesh A, Mehrabian MA. Optimization of a lithium bromide–water solar absorption cooling system with evacuated tube collectors using the genetic algorithm. Energy and Buildings. 2014;85:427-35. [13] López-Villada J, Ayou DS, Bruno JC, Coronas A. Modelling, simulation and analysis of solar absorption power-cooling systems. International Journal of Refrigeration. 2014;39:125-36. [14] Domínguez-Inzunza LA, Hernández-Magallanes JA, Sandoval-Reyes M, Rivera W. Comparison of the performance of singleeffect, half-effect, double-effect in series and inverse and triple-effect absorption cooling systems operating with the NH3– LiNO3 mixture. Applied Thermal Engineering. 2014;66:612-20. [15] Li Z, Ye X, Liu J. Optimal temperature of collector for solar double effect LiBr/H2O absorption cooling system in subtropical city based on a year round meteorological data. Applied Thermal Engineering. 2014;69:19-28. [16] Avanessian T, Ameri M. Energy, exergy, and economic analysis of single and double effect LiBr-H2O absorption chillers. Energy and Buildings. 2014;73:26-36. [17] Said SAM, El-Shaarawi MAI, Siddiqui MU. Analysis of a solar powered absorption system. Energy Conversion and Management. 2015;97:243-52. [18] Fong KF, Chow TT, Lee CK, Lin Z, Chan LS. Solar hybrid cooling system for high-tech offices in subtropical climate-Radiant cooling by absorption refrigeration and desiccant dehumidification. Energy Conversion and Management. 2011;52:2883-94. [19] Ketjoy N, yongphayoon R, Mansiri K. Performance Evaluation of 35 kW LiBr-H2O Solar Absorption Cooling System in Thailand. Energy Procedia. 2013;34:198-210. [20] Lu ZS, Wang RZ, Xia ZZ, Lu XR, Yang CB, Ma YC, et al. Study of a novel solar adsorption cooling system and a solar absorption cooling system with new CPC collectors. Renewable Energy. 2013;50:299-306. [21] Ali AHH, Noeres P, Pollerberg C. Performance assessment of an integrated free cooling and solar powered single-effect lithium bromide-water absorption chiller. Solar Energy. 2008;82:1021-30. [22] Hang Y, Qu M, Winston R, Jiang L, Widyolar B, Poiry H. Experimental based energy performance analysis and life cycle assessment for solar absorption cooling system at University of Californian, Merced. Energy and Buildings. 2014;82:746-57. [23] Yin YL, Zhai XQ, Wang RZ. Experimental investigation and performance analysis of a mini-type solar absorption cooling system. Applied Thermal Engineering. 2013;59:267-77. [24] Lazzarin RM, Romagnoni P, Casasola L. Two years of operation of a large solar cooling plant. International Journal of Refrigeration. 1993;16:185-90. [25] Xu SM, Huang XD, Du R. An investigation of the solar powered absorption refrigeration system with advanced energy storage technology. Solar Energy. 2011;85:1794-804. [26] Venegas M, Rodríguez-Hidalgo MC, Salgado R, Lecuona A, Rodríguez P, Gutiérrez G. Experimental diagnosis of the influence of operational variables on the performance of a solar absorption cooling system. Applied Energy. 2011;88:1447-54. [27] Al-Ugla AA, El-Shaarawi MAI, Said SAM. Alternative designs for a 24-hours operating solar-powered LiBr-water absorption air-conditioning technology. International Journal of Refrigeration. 2015;53:90-100. [28] Prasartkaew B. Performance Test of a Small Size LiBr-H2O Absorption Chiller. Energy Procedia. 2014;56:487-97. [29] Calise F, Palombo A, Vanoli L. Maximization of primary energy savings of solar heating and cooling systems by transient simulations and computer design of experiments. Applied Energy. 2010;87:524-40. [30] Noro M, Lazzarin RM. Solar cooling between thermal and photovoltaic: An energy and economic comparative study in the Mediterranean conditions. Energy. 2014;73:453-64.
Page 14 of 24
Wind velocity sensor
Solar beam radiation meter
Blower coil
Data acquisition system
Blower coil
Blower coil
Meeting room (102 m2) T15
T14
T13
Supplement water tank
Make up water Supplement water tank
T12 T4
T3
Hot water storage tank
T8
T6
TX-23
T11
PTC2 T5
T2
T1
Chilled water tank
Make up water
Pollution Discharge vavle
Hot water pump
LiBr absorption cooling chiller
T7 Pollution Chilled water pump discharge vavle T10 Cooling tower
T9 Cooling water pump
PTC1
Pump
Pollution discharge vavle
Pollution discharge vavle
Valve
Fig. 1 – Scheme diagram of solar powered absorption cooling system.
Page 15 of 24
Fig. 2 –Main parts of the solar absorption cooling system.
Sun
Qp,loss,1
Ei Qtc,ins PTC array
Pipeline
Hot water Storage tank
pipeline
The end of air-condition
Qa,room
Pipeline
Qp,a,3
Absorption cooling chiller
Qt,loss
Qp,loss,2
Cooling tower
Qe
Qc
Fig. 3 – Heat flows diagram of the solar absorption cooling system.
Page 16 of 24
0.8
35
0.9
0.7
30
0.8
0.4 15
0.3 te,ins
0.2
10
Ptc,ins Ib
0.1 0.0 12.5 12:30
13.0 13:00
13.5 13:30
14.0 14:00
14.5 14:30
0.5 0.4 0.3
5
15.0 15:00
0.6
-2
20
Ptc,ins (kW)
te,ins
0.7
25
0.5
Ib (kW m )
0.6
0.2
0 15.5 15:30
0.1
Time (h) Time (hh:mm) Time (hh:mm)
Fig. 4 – Test results of instantaneous efficiency and instantaneous power of PTC array.
0.7
35
0.9
0.6
30
0.8
0.5
25
0.4
20
0.3
15
0.2
10
te,ins,s
-2
0.5
Ib (kW m )
0.6
0.4 0.3 0.2
Ptc,ins,s Ib
0.1
0.7
Ptc,ins,s (kW)
te,ins,s
Fig. 4 – Test results of instantaneous efficiency and instantaneous power of PTC array.
5
0.0 0 12.5 12:30 13.0 13:00 13.5 13:30 14.0 14:00 14.5 14:30 15.0 15:00 15.5 15:30
0.1 0.0
Time Time (h)(hh:mm) Fig. 5 – Thermal efficiency and power of heat collection system.
Fig. 5 – Thermal efficiency and power of heat collection system. 1.0
5
0.8
4
0.4
3
0.3
2
0.2 1
0.1 0.0
8:00 8
10:00 10
12:00 12
14:00 14
Time (hh:mm) Time (h)
(a) 2014-05-11
16:00 16
0
18:00 18
2
0.5
8 Ib
vw
7 6 5
0.6
4 0.4
3 2
0.2 0.0
8:00 8
vw (m/s)
vw
6
2
Ib (kW/m )
0.6
Ib
Ib (kW/m )
0.7
vw (m/s)
0.8
1 10:00 10
12:00 12
14:00 14
16:00 16
0
18:00 18
Time (hh:mm) Time (h)
(b) 2014-05-14
Page 17 of 24
Fig. 6 – Solar beam radiation and wind velocity vs. time. T5 T7 T9
70 60
T6 T8 T10
70 60 50
T ( C)
40
o
o
T( C)
50
30
30
20
20
10
12:54
13:54
14:54
15:54
10
16:54
T5 T7 T9
40
T6 T9 T10
12:19 13:19 14:19 15:19 16:19 17:19 18:19
Time (hh:mm)
Time (hh:mm)
(a) 2014-05-11
(b) 2014-05-14
0.6
16 r
r
14
Pr
0.5
1.0
6
0.2
0.6
r
8
Pr (kW)
10
0.3
4 0.1 0.0
0.2
2 13 13:00
14 14:00
13:00
14:00
15 15:00
16 16:00
17 17:00
0 18 18:00
17:00
18:00
0.4
0.0 12:00 12 13:00 13 14:00 14 15:00 15 16:00 16 17:00 17 18:00 18 19:00 19
Time (h) Time (hh:mm) 15:00
Time (h) Time (hh:mm)
16:00
(a) 2014-05-11
Time Time (hh:mm) (hh:mm)
(b) 2014-05-14
Fig. 8 – The refrigeration coefficient and refrigeration power of the absorption cooling chiller. 28
24 23
26
22
24
T ( C)
20
o
T ( C)
21 o
r
0.4
Pr
0.8
12
22 20 18 16 14 12 10 8 6 4 2 0
Pr (kW)
Fig. 7 – The inlet and outlet temperatures of heating, chilled and cooling of chiller vs. time.
19 18
Tr1 Tr2 Ta
17 16
12:54
13:54
14:54
15:54
16:54
22 Tr1 Tr2 Ta
20 18
12:19 13:19 14:19 15:19 16:19 17:19 18:19
Time (hh:mm)
Time (hh:mm)
(a) 2014-05-11
(b) 2014-05-14
Page 18 of 24
Fig. 9 – The variation of indoor temperatures during refrigerating. 50 45
35
o
T ( C)
40
30
T5 T6 T7 T8
25 20
15:48 15.8 16:00 16.0 16:12 16.2 16:24 16.4 16:36 16.6 16:48 16.8 17:00 17.0 17:12 17.2 17:24 17.4
Time (h) Time (hh:mm) Fig. 10 – The running temperatures of absorption chiller when it was under the heating mode (Mar. 29). Fig. 10 – The running temperatures of absorption chiller when it was under the heating mode (Mar. 29). 28 26
o
T ( C)
24 22 20
Tr1 Tr3
Tr2 Ta
18 15:48 15.8 16:00 16.0 16:12 16.2 16:24 16.4 16:36 16.6 16:48 16.8 17:00 17.0 17:12 17.2 17:24 17.4
(hh:mm) Time (h) Time (hh:mm) Fig. 11 –running The variations of room temperatures during theitheating mode 29).mode (Mar. 29). Fig. 10 – The temperatures of absorption chiller when was under the(Mar. heating
Fig. 11 – The variations of room temperatures during the heating mode (Mar. 29). 22 20
r
Pr
0.8
18
0.7
16
0.6
14 12
0.5
r
10
0.4
8
0.3
6
0.2
4 2
0.1 0.0 46
Pr (kW)
0.9
0 48
50
52
54
56
58
60
62
64
66
68
70
72
74
o
T5 ( C)
Page 19 of 24
Fig. 12 – Instantaneous refrigeration coefficient vs. running temperatures. Fig. 12 – Refrigeration coefficient vs. running temperatures.
31
0.75 T9,
30
T10,
r
0.70
29
0.65
28
0.60
26
0.55
25
0.50
24 23
0.45
22
0.40
21 20 13:36 13.6
r
o
T ( C)
27
13:48 13.8
14:00 14.0
14:12 14.2
14:24 14.4
14:36 14.6
14:48 14.8
0.35
15:00 15.0
Time (h) Time (hh:mm)
Fig. 13 – Cooling performance influent by cooling water temperature.
Fig. 13 – Cooling performance influent by cooling water temperature. Table 1 – Instructions for the temperature probes in Fig. 1. Temperature Measurement explain probe T1
Inlet temperature of PTC1
T2
Outlet temperature of PTC1
T3
Inlet temperature of PTC2
T4
Outlet temperature of PTC2
T5
Hot water inlet temperature of chiller
T6
Hot water outlet temperature of chiller
T7
Chilled water inlet temperature of chiller
T8
Chilled water outlet temperature of chiller
T9
Cooling water inlet temperature of chiller
T10
Cooling water outlet temperature of chiller
T11
Outlet temperature of hot water tank
T12
Inlet temperature of hot water tank
T13, T14
The indoor temperature of meeting room
T15
The outdoor temperature of meeting room
Page 20 of 24
Table 2- Main measurements and instruments accuracy of the monitoring system. Instrument
Model
Range
Accuracy
Application scope
Maximum relative error
Maximum absolute error
Uncertainty (B class)
Pyranometer
Kipp & Zonen CMP-6
0-2000 2 (W/m )
±5%
0-1000 2 (W/m )
±10%
±100W/m2
57.7348 2 W/m
Thermocouples
T
-200 to 350 (℃)
±0.4%
0-150 (℃)
±0.93%
±1.4℃
0.8083℃
Wind speed transducer
EC-9S
070(m/s)
±0.4%
0-10 (m/s)
±2.8%
±0.28m/s
0.1617 m/s
Electromagnetic flow meter
KROHNE OPTIFLUX 5300
DN 25; 012 (m/s)
±0.15%
0-5 (m/s)
±0.36%
±0.018m/s
0.0104 m/s
Pressure transducer
YOKOGAWA EJA430E
0.14-16 (MPa)
±0.055%
0-2 (MPa)
±0.44%
±0.0088MPa
0.0051 MPa
Table 3 – PTC array parameters. Parameter
Value
Aperture area
56 m
Aperture width
2.5 m
Orientation
North-South (ψ = 0°)
Length of PTC
26 m
Focal distance of parabolic trough
1.1 m
The width of focal spot
5 cm
PTC efficiency values at different temperature
77.5%(30°C), 68.4%(50°C), 57.3%(70°C), 38.4%(90°C)
Inner diameter of metal pipe of receiver
4 cm
Outer diameter of glass pipe of receiver
11 cm
2
Page 21 of 24
Table 4 – Parameters of absorption chiller. Parameter
Value
Model
TX-23
Working pair
Lithium Bromide-Water
Ambient temperature
28°C -36.0°C
hot water temperature at chiller inlet
50°C -90.0°C
hot water temperature at chiller outlet
49°C -80.0°C
Outlet temperature of chilled water
10.0°C-14.0°C
Inlet temperature of chilled water
15.0°C-22.0°C
Cooling water temperature at chiller inlet
20.7°C-27.8°C
Cooling water temperature at chiller outlet
23.5°C-29.3°C
Power dissipation
2.3 kW
Refrigeration capacity
23 kW
Chilled water rate
4.0 m3/h
Hot water rate
5.7 m3/h
Table 5 – Parameters of cooling tower, tank and pumps. Parameter
Value
Model
BLT-10
Wind rate
10.5 m /h
Cooling water rate
10 m /h
Dynamo power
0.75 kW
Capacity of tank
1m3
Insulation thickness of tank
3 cm
Pump of PTC
0.33 kW
Hot water pump
0.78 kW
Chilled water pump
0.78 kW
Cooling water pump
0.98 kW
Comment [A1]: Author: There are two different Table 5 captions were provided in the manuscript, and this has been retained. Please check and confirm it is correct.
3
3
Page 22 of 24
Table 6 – Parameters of blower coil. Parameter
Value
Model
EKCW800KT
Wind rate
1360 m /h
Cooling power
7200 W
Input power
130 W
3
Table 7 – Parameter of pipes and the energy losses of the pipelines. Pipeline type
Pipeline between PTC and hot water storage tank Pipeline between hot water storage tank and chiller Pipeline between chiller and blower coil
Inner tube diameter
Outer tube diameter
Insulation thickness
Pipe length
Total energy losses (kW)
4 cm
12.5 cm
4 cm
78 m
3.90–6.20
4cm
12.5 cm
4 cm
18 m
0.90–1.40
4cm
10 cm
3 cm
55 m
0.27–0.55
Note
Comment [A2]: Author: There are two different Table 7 captions were provided in the manuscript, and this has been retained. Please check and confirm it is correct.
60–90 ºC (heat loss) 60–90 ºC (heat loss) 7–16 ºC (cooling loss)
Table 8 – The average coefficient of performance for the refrigeration system. Date (2014)
Weather condition
Ei,tot (MJ)
Qr (MJ)
ηr,av
COPs,av
Apr. 4
Cloudy and windy
448
54
0.18
0.12
May 10
Cloudy and windy
643
84
0.21
0.13
May 11
Cloudy and windy
549
58
0.17
0.11
May 13
Sunny and windy
1300
220
0.51
0.17
May 14
Sunny and windy
1200
170
0.45
0.14
May 15
Sunny and windy
1122
234
0.57
0.21
May 16
Sunny and windy
1120
267
0.58
0.23
May 17
Sunny and windy
1280
284
0.51
0.22
May 18
Sunny and windy
1262
313
0.55
0.25
Page 23 of 24
May 19
Sunny and windy
877
288
0.59
0.23
May 20
Cloudy and windy
690
158
0.51
0.21
May 21
Sunny and windy
967
270
0.57
0.27
May 22
Sunny and windy
1110
262
0.47
0.24
May 24
Sunny and windy
1084
234
0.50
0.22
Table 9
– The gained heat of system in the heating mode.
Date (2014)
Weather condition
Ei,tot (MJ)
Qh (MJ)
Average gained heat rate
Mar. 2
Sunny and windy
1032
412
0.399
Mar. 29
Cloudy and windy
423
164
0.387
May 12
Cloudy and windy
197
81
0.411
Page 24 of 24
S0140-7007(16)30234-1 http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.07.023 JIJR 3396
To appear in:
International Journal of Refrigeration
Received date: Revised date: Accepted date:
7-1-2015 11-7-2016 30-7-2016
Please cite this article as: Ming Li, Chengmu Xu, Reda Hassanien Emam Hassanien, Yongfeng Xu, Binwei Zhuang, Experimental investigation on the performance of a solar powered lithium bromide-water absorption cooling system, International Journal of Refrigeration (2016), http://dx.doi.org/doi: 10.1016/j.ijrefrig.2016.07.023. This is a PDF file of an unedited manuscript that has been accepted for publication. As a service to our customers we are providing this early version of the manuscript. The manuscript will undergo copyediting, typesetting, and review of the resulting proof before it is published in its final form. Please note that during the production process errors may be discovered which could affect the content, and all legal disclaimers that apply to the journal pertain.
Experimental investigation on the performance of a solar powered lithium bromide-water absorption cooling system
Ming Li1,1, Chengmu Xu1, Reda Hassanien Emam Hassanien1,2, Yongfeng Xu1,3, Binwei Zhuang1. 1
Solar Energy Research Institute, Yunnan Normal University, Kunming 650500, China.
2
Agricultural Engineering Department, Faculty of Agriculture, Cairo University, Cairo 12613, Egypt.
3
Zhejiang Solar Energy Product Quality Inspection Center, Haining, Zhejiang 314416, China.
Highlights
The cooling and space heating performance of the system was experimental investigated.
The cooling performance of the system influent by running temperature was analyzed.
The cooling performance of the system influent by cooling temperature was analyzed.
The improvement methods for the system were analyzed and discussed.
Abstract: The Performance of solar cooling absorption system needs further research, due to its poor coefficient of performance (COP).Therefore; this study investigated the performance of a 23 kW solar powered single-effect lithium bromide-water (LiBr-H2O) absorption cooling system. Furthermore, the space heating mode was also investigated and the improvement methods were analyzed and discussed. The cooling system was driven by a parabolic trough collector of 56 m2 aperture area and used for cooling a 102 m2 meeting room. Research results revealed that the chiller’s maximum instantaneous refrigeration coefficient (chiller efficiency) could be up to 0.6. Most of the time, in sunny and clear sky days the daily solar heat fraction ranged from 0.33 to 0.41 and the collectors field efficiency ranged from 0.35 to 0.45. At the same time, chiller efficiency was varied from 0.25 to 0.7 and the daily cooling COP was varied from 0.11 to 0.27, respectively. Keywords: Solar cooling; Single-effect absorption chiller; Lithium Bromide-water; Parabolic trough solar collector (PTC); Cooling performance
Nomenclature Symbols
1
Corresponding author. Tel./fax: +86 871 65517266.
E-mail address: [email protected] (M. Li).
Page 1 of 24
Ac
collection area of PTC, m2
cp
specific heat, J kg-1 K-1
COPs,av
Average COP of the whole system.
Ei
input solar radiation into PTC, W
Ei,tot
total input solar radiation energy into PTC, MJ
Ib
solar beam radiation, W m-2
mF
mass flow rate of water which flow passing by PTC array, kg s-1
mF,g
mass flow rate of water flow passing by the chiller’s generator, kg
mF,c
mass flow rates chilled water, kg s-1
mF,e
mass flow rates of cooling water, kg s-1
mw,t
water mass in the tank, kg
Pte,ins
instantaneous thermal power of PTC, W
Pte,ins,s
instantaneous power of heat collection system, W
Pr
the gained refrigeration quantity from the absorption chiller system, W
Pg
consumption thermal power of generator , W
Pe
absorption power from chilled water of water evaporate, W
Pc
output power into cooling water by absorber and condenser, W
Qtc,ins
the received useful solar radiation energy, W
Qp,loss
pipeline heat losses, W
Qt,loss
heat losses of the hot storage tank, W
Qr
total refrigeration quantity, MJ
Qh
the gained heat of the system, MJ
T1 – T15
the temperatures of probes in Fig. 1, ºC
Ta
corridor temperature (T15 in Fig. 1), ºC
Tr
indoor (meeting room) temperatures, ºC
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Greek symbols ΔT
temperature difference, ºC
ηte,ins
the instantaneous thermal efficiency
ηte,ins,s
the instantaneous thermal efficiency of heat collection system
ηr
thermodynamics coefficient
ηoel
the optical end loss ratio of PTC
ηr,av
the average refrigeration efficiency
θ
incidence angle, °
Abbreviations COP
coefficient of performance
LiBr-H2O
lithium bromide-water
PTC
parabolic trough solar collector(s)
1. Introduction The global climate changes resulted in an increasing of energy demand for traditional air conditioning and heating systems in buildings which, consumes more fossil fuel meanwhile, increasing the carbon emissions. Therefore, using solar energy for air conditioning becomes one of the promising approaches to reduce energy consumptions and negative environmental impacts from buildings. There are two main types of solar cooling namely solar-thermal-driven and solar-PV-driven air conditioning technologies. Solar-PV-driven system uses a conventional vapor compression air conditioning cycle in which the electrical input is provided by solar PV panels. The solar-thermal-driven utilizes solar thermal energy to power the generator of an absorption refrigeration systems and it has three cooling technologies such as, solar absorption refrigeration, solar adsorption refrigeration and solar ejector refrigeration[1, 2]. However, solar absorption refrigeration is the most matured technologies and it has been studied more extensively than other systems. The main advantage of the solar absorption cooling technology is that the coefficient of performance (COP) is higher than that of other thermally operated cycles [3]. Furthermore, in a comparison between the different solar electric, solar thermal from the point of view of energy efficiency and economic feasibility, solar electric and thermo-mechanical systems appear to be more expensive than thermal sorption systems. Absorption and adsorption are comparable in terms that adsorption chillers are more expensive and bulkier than absorption chillers. The total cost of a single-effect LiBr-H2O absorption system is estimated to be the lowest [4]. So due to relatively simple configuration and low requirements for heat sources, a numerous studies of solar absorption refrigeration system are using single-effect LiBr-H2O absorption chillers, recently [5-10]. Furthermore, Double-effect LiBr-H2O absorption chiller systems coefficients of performance are almost twice higher than those obtained with single-effect systems. with Double-effect systems it is possible to obtain
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coefficients of performance as high as 1.12 at condenser temperatures of 30 °C but they need generator temperatures higher than 140 °C to reach evaporator temperatures as low as -5 °C [1115]. But there are some high requirements for heat sources maybe more than 120°C. So Double-effect LiBr-H2O system is sophisticated construct and high investment and operation cost. Nevertheless, Avanessian and Ameri reported that the CO2 emission of the single-effect system was respectively about 1.9 and 1.7 times higher than direct-fired and hot-water double-effect [16]. The efficiency enhancement of the solar absorption cooling systems components is essential to increase the COP of the whole system [17, 18]. The performance evaluation of a 35 kW LiBr-H2O absorption machine driven by 72 m2 heat pipe evacuated tube collector with a gas backup system showed that the actual average COP of the system was 0.33 while the maximum and minimum values were 0.50 and 0.17 respectively. These results were obtained based on the average collector efficiency and the solar fraction of 0.55 and at an ambient temperature of 32°C[19]. Lu et al.[20] investigated one two-phase thermo-syphon silica gel-water solar adsorption chiller and LiBr-H2O absorption chiller with new medium Compound Parabolic Concentrator (CPC) solar collectors. Results revealed that the efficiency of the medium temperature evacuated-tube CPC solar collector can reach 0.5 when the hot water temperature is 125°C. The average solar COPs of absorption system was 0.19 [20]. On the other hand, Ali et al. [21] investigated the Performance assessment of an integrated free cooling and solar powered single-effect LiBr-H2O absorption chiller in Oberhausen (Germany). The plant includes 35.17 kW cooling absorption chiller, 108 m2 evacuated tube collectors, a 6.8 m3 hot water tank, a 1.5 m3 cold water tank and a 134 kW cooling tower. The results illustrated that the chiller efficiency in some cooling months could be up to 70% while it was about 25% during the 5 years period of the plant operation; the monthly average efficiency value of collectors varies from 34.1% up to 41.8% and the five-year average value is about 28.3%. Hang et al. [22] investigated the energy performance analysis of solar absorption cooling system, a 23 kW double-effect absorption chiller driven by a 54 m2 External Compound Parabolic Concentrator (XCPC) solar collectors. The results showed that the daily average collector efficiency changed between 36% and 39%. The average coefficient of performance (COP) of the LiBr absorption chiller was between 0.91 and 1.02 with an average value of 1.0, and the daily solar COP was approximately at 0.374. On other study, the performance analysis of a mini-type solar-powered absorption cooling system with a cooling capacity of 8 kW, solar collector’s area of 96 m2 and a water storage tank of 3 m3 was conducted. When LiBr-H2O used as the working pairs of the chiller, the experimental results revealed that the average values of predicted mean vote and predicted percentage of dissatisfied of the test room were 0.22 and 5.89, respectively, and the power consumption was reduced by 43.5%. Meanwhile, the theoretical model predicted that the solar radiation intensity has a great impact on the performance of the solar powered absorption cooling system compared with the ambient temperature[23]. Lazzarin et al. [24]reported their field test results for a solar cooling plant and gave the performance of a typical day in terms of energy collected and temperature of the solar collector. They indicated that both energy collected and temperature of the solar collector were significantly various during the day. Due to the solar input constantly varies. Subsequently, the performance of the system may deteriorate and the solar powered refrigeration system will not be able to work consistently[25]. Venegas et al.[26] investigated the influence of operational variables on the performance of a solar absorption cooling system. Results showed that the most important variables which effected on the daily cooling energy produced and the daily averaged solar COP are the amount of
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the collected solar energy, the wind velocity and wind direction. On the other hand, it was reported that the optimum mass flow rates of hot water passing through the generator on a double-effect absorption chiller with 100 t of cooling capacity and evacuated tube collector have an important role on reducing the auxiliary energy[12] .Subsequently, three alterative designs (heat storage, cold storage, and refrigerant storage) for 24-hour-operating solar-powered LiBr-H2O absorption air-conditioning systems were analyzed. The results indicate that continuously operating solar-powered LiBr-H2O absorption with refrigerant storage was the most suitable alternative design for a 24-hour cooling effect[27]. It was demonstrated that, to obtain the high coefficient of performance for a small size absorption chiller which was developed from an old out-of-order commercial chiller, this chiller should be operated at 85°C of hot water temperature supplied to the generator[28]. Therefore, Calise et al. [29]simulated different solar cooling systems to find out the appropriate operation parameters for maximizing the COP of solar cooling systems. However, there are many challenges are remaining for further studied to enhance the COP of the absorption refrigeration technology. This study investigates the experimental performance of a single-effect lithium bromide absorption refrigeration system driven by PTC for air conditioning. Furthermore, it discusses and analyzes appropriate methods for improving the cooling performance.
2. System description This experiment has been conducted in Kunming, China. The collector was a parabolic trough solar collector (PTC). Fig. 1 shows the schematic diagram for the main parts of the refrigeration experimental system which includes PTC array, hot water storage tank (with a supplemental water tank), single-effect lithium bromide absorption chiller (TX-23), cooling tower and blower coils installed in the meeting room. Fig. 1 also illustrates the positions of temperature test points. Table 1 shows the sensors of temperatures and its instruction. Fig. 2 shows the main parts of the refrigeration experimental system. The Main measurements and instruments accuracy of the monitoring system are shown in table 2. The parameters and heat efficiency values at different temperature of PTC array were shown in table 3. The heat receiver, glass-metal evacuated absorber tube, was the core component of PTC and the specifications were shown in table 2 too. The PTC array is made up with two PTCs in series and the array space between two PTCs was 5.6 m. Table 4 shows some parameters of refrigeration chiller. The capacity of hot water storage tank was 1 m3. Table 5 shows the cooling tower parameters, the same type of cooling water has been used in the cooling tower. The area of end air-conditioning room was 102 m2 and 3 Brower coils installed on the top of the meeting room (relevant parameters are shown in Table 6). The room is used to conference room which can accommodate more than 20 persons. The high of the room is 3.5m. And the length and the width are 13.6 m and 7.5 m, respectively. There are two windows with 10m2 transparent fiberglass in the north wall. The south wall nearby the corridor with two glass doors and the area of each door is 3 m2. And then, the last two walls are made up with reinforced concrete with 10cm thickness.
3. Energy conversion Analysis
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Fig. 3 shows the heat flow diagram of the absorption chiller, its main energy conversion processes are as follows: (1) The input solar radiation into PTC: E i I b Ac cos
(1) Where, Ib, Ac and θ are the solar beam radiation, the collection area and the incident angle of sun ray, respectively. (2) The useful energy transformed from the received solar radiation Qtc,ins. According to Fig. 1, Q tc , ins m F c p [(T2 T1 ) (T4 T3 )]
(2)
Where, mF is the mass flow rate of water flowing in the receiver of PTC array, and cp is the specific heat capacity at constant pressure. (3) The pipeline heat loss between PTC and hot water storage tank Qp,loss,1: Q p , loss ,1 m F c p [(T11 T1 ) (T2 T3 ) (T4 T12 )]
(3)
(4) The calculation of the heat loss of hot water storage tank Qt,loss is relatively complicated. The The heat loss of hot water tank was tested in accordance with ISO 9459 and EN12976. Before the experiment, the water in the tank was heated by PTC and the temperature was not low than 60°C. In order to ensure the temperature consistency in the tank and avoid tank thermal stratification, an external water loop was adopted. The water was pumped by hot water pump out of the tank in the bottom and back into the tank in the top and then the hot water was circulated through the external water pump. The pump and the circulation pipes were wrapped with insulation material, the thickness of which was no less than 3cm. There temperature sensors was arranged in equal intervals in the vertical direction inside of the hot water tank. When the values of there sensors were close, there are no thermal stratification in the tank. And then the circulation pump must be shut down. The hot water temperature was 63.6°C. The water was placed for 15 hours without any interfere. And then, the circulation pump was turn on once again and the hot water was circulated until the temperatures of sensors were the same nearly. The temperature was 49.8°C. In real experimental test we can use the simple formula to calculate it as follows: Pt, loss , av
m w ,t c p T t
(4)
Where, mw,t is the water mass in the tank, Δt is the continue time of experiment, and ΔT is the temperature different of water with no thermal stratification in the tank within the time of Δt. When, Δt→0, it can be gained the instantaneous heat loss power.
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(5) The pipeline heat loss between hot water storage tank and chiller: Q p , loss , 2 2 m F , g c p (T11 T5 )
(5)
Where, mF,g is the mass flow rate of water flow passing by the chiller’s generator. (6) The heat absorbed from chilled water in the evaporator of refrigeration chiller, Qp,loss,2 (this issue will be analyzed in the next section). (7) The output heat from absorber and condenser to cooling water, this part of heat can be equivalent to the output heat Qc from cooling tower. (8) Qp,a,3 is the heat absorbed of pipeline between chiller and meeting room which comes from the out surface of pipeline. (9) Qa,room is the heat absorbed of blower coils which comes from the air in the meeting room.
4. Thermal efficiency of heat collection system 4.1. Thermal efficiency of PTC array The instantaneous thermal power and instantaneous thermal efficiency are respectively as follows: Pte , ins m F c p [(T 2 T1 ) (T 4 T3 )] te , ins
Pte , ins I b Ac cos
(6) (7)
Because of the roof orientation factor the PTC array could shade by adjacent buildings after 4:30. Therefore, the experiments could only be conducted before 4:30. The experiments have been also conducted under the windy weather conditions, the tested peak wind velocity reached to 8 m s -1, which was largely influence to the thermal efficiency of PTC array under such weather condition. Because the focal line was frequently move out of the absorber tube under such high wind velocity. Fig. 4 gives the test results of thermal efficiency of PTC array. Obviously, when the irradiance changed from 0.80 to 0.90 kW m-2 the instantaneous thermal efficiency ranged from 0.50 to 0.65. at the same time, the system instantaneous thermal power was changed between 24 and 27 kW. In the calculations, the mass flow rate of water mF was tested as 0.602 kg s-1. 4.2. Energy losses from the pipelines Here, the test of energy losses of the pipeline between PTC and hot water storage tank, energy losses of the pipeline between hot water storage tank and chiller, and energy losses of the pipeline between chiller and meeting room were conducted.
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The heat conduction coefficient of thermal insulation layer of pipeline was between 0.037-0.040 W mK-1 and the out surface of the thermal insulation was wrapped with aluminum alloy layer. And the three types of pipelines heat loss are shown in Table 7. It is worth pointing out that because of the temperature of chilled water inside the tube was lower than that of the outside, the heat was absorbed from outside to inside of tubes. Therefore, the energy losses of the pipeline between chiller and meeting room were cooling losses, rather than heat losses. 1
4.3. Heat losses of hot water storage tank The heat loss of water storage tank was test. At first, the tank was filled with 688kg hot water. And 15 hours later, the temperature of water has been decreased from 63.6°C to 49.8 °C. It can be estimated that the average heat losses power was about 640 W. Obviously, the average power of heat losses increased with the increase of hot water temperature. 4.4. Thermal efficiency of heat collection system When the pipeline heat losses and heat losses of hot water tank are considered, the instantaneous power of heat collection system can be expressed as: Ptc , ins , s Ptc ,ins Pp ,loss Pt,loss
(8)
While the instantaneous thermal efficiency of heat collection system can be expressed as: te , ins , s
Ptc , ins Pp , loss Pt,loss I b Ac cos
(9)
Where, Pp,,loss and Pt,loss are pipelines heat losses and hot water tank heat losses, respectively. Thus when the solar beam radiation changed between 0.80 and 0.90 kW m-2, the instantaneous thermal efficiency and the instantaneous power ranged from 0.32 to 0.42 and from 17 to 21 kW, respectively. (Fig. 5).
5. Performance of the refrigeration system This experiment was conducted mainly to test and investigate the performance of the refrigeration system. During the experiment many parameters such as solar beam radiation, wind velocity, ambient temperature, inlet and outlet temperature of PTC, hot water storage tank, heating, chilled water, and cooling water were tested. 5.1. Energy analysis of refrigeration process for cooling chiller The thermodynamics coefficient (ηr) was considered as the economy evaluation which defined as r Pe Pg
(10)
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Where, Pe is the absorption power from chilled water of water evaporates, and Pg is the consumption of heat energy. According to Fig. 1, the consumption thermal power of generator is as follows: Pg m F , g c p (T 6 T5 )
(11)
The absorption power from chilled water of water evaporates is: Pe m F , e c p (T 7 T8 )
(12)
The output power into cooling water by absorber and condenser is as the following: Pc m F , c c p (T10 T9 )
(13)
In the three equations above, mF,g, mF,c and mF,e are mass flow rates of hot water, chilled water and cooling water, respectively. T5 - T10 are the inlet and outlet temperatures of hot water, chilled water and cooling water (see Fig.1 and Table 1). According to the analysis above, if ignored the power consumed by circulating water pump, the cycle refrigeration coefficient of the chiller (chiller efficiency or COP of the chiller) can be express as： r
Pe Pg
m F , e c p ( T 7 T8 ) m F , g c p (T 6 T5 )
m F , e ( T 7 T8 ) m F , g (T 6 T5 )
(14)
5.2. Experimental results The cooling performance of the absorption chiller for ten days of experiment was conducted. The experiment was conducted from April to May, 2014 at Yunnan Yi Tong Solar Science and Technology Co. Ltd (in Kunming, China). At first the water in tank must be heated by PTC. When the temperature of hot water in the tank reached more than 65°C, the refrigeration chiller was turned on by manual. In order to understand the chiller performance under different operating temperatures, the different water temperature is adopted to driven chiller and the refrigeration performance can be got and recorded. But in order to protect refrigeration chiller, the test experiment must be shut down when the hot water temperature in the tank descended to around 40 °C. All of the components are controlled by manual. During experiment, the test value of mF, mF,g and mF,c are 0.602 kg s-1, 1.36 kg s-1 and 0.90 kg s-1 respectively. The experimental results are shown table 8 and Figs. 6, 7, 8 and 9 are shown the changing curves of system performance parameters along with time. Where, the figures show the experimental results in two days and Table 8 gives the experimental results for refrigeration performance during 14 days.
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According to Fig. 6, during the experiment, the wind speed is high in most sunny and cloud days and the maximum wind velocity vw reached to 8 m s-1 (generally it ranged between 2 and 5 m s-1), which had a huge impact on the thermal efficiency of PTC system. Fig. 7 shows that when the operating temperature of refrigeration chiller (the temperature of hot water) increased from 39 ºC to 80ºC, the lowest temperature of chilled water could reach to 9 ºC. At the same time, the temperature of cooling water changed between 20 ºC and 35 ºC. The cooling process was disconnected. At the beginning of the experiment, the temperatures of hot water and chilled water decreased rapidly. After chiller operating steadily, the system performance parameters changed stably along with time. The higher the inlet temperature of hot water, the lower the outlet temperature of the chilled water is. Fig. 8 shows the changing curves of refrigeration coefficient and refrigeration power of chiller along with time. It shows that the refrigeration coefficient ηr increased along with the refrigeration quantity Pr increase of chiller. Fig. 9 shows that in the first thirty minutes after refrigeration chiller operating, the indoor temperature of meeting room decreased sharply. After reaching a certain value, the indoor temperature began to gradually stabilize. Sometimes, the temperature fluctuated slightly. The indoor temperature increased along with the ambient temperature. Table 8 shows the average value of performance coefficient of the refrigeration system during each experimental day. Where, Ei,tot is the total incident solar radiation energy (the total input solar radiation energy into PTC), Qr is the total refrigeration quantity, ηr,av is the average refrigeration efficiency and COPs,av is the average COP of the refrigeration system. The total incident solar radiation energy Ei,tot is given as E i , tot
t s , st
t s ,bt
Ac I b cos (1 oel )dt S
(15)
Where, ts,bt and ts,st are the time of experiment starting and finishing, respectively, ηoel is the optical end loss ratio of PTC (Xu et al., 2014). The total refrigeration quantity is given as Qr
t r , st
t r ,bt
m F , e c p (T7 T8 )dt
(16)
Where, tr,bt and tr,st are the time of refrigeration chiller operating and stopping, respectively. The average refrigeration efficiency expressed as r , av
Qr
t r , st t r ,bt
m F , g c p (T5 T6 )dt
(17)
COPs,av is shown as
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C O Ps , av
Qr
(18)
E i , tot
Table 8 illustrates that the average refrigeration efficiency (ηr,av) and the average COP (COPs,av) of the whole system were relatively low under cloudy weather conditions. In such weather condition, the solar radiation was relatively small. So the hot water could not be heated to high temperature. In addition, when the sun covered by cloud, the pump was running with fixed frequency all the time which resulted in more pipeline heat losses. Therefore, the average refrigeration efficiency and the average COP of the system were relatively low. In contrast, under clear weather conditions, the refrigeration efficiency and system COP were high. The experiment results revealed that the average refrigeration efficiency of chiller changed between 0.17 to 0.60, and the refrigeration coefficient of the whole system ranged from 0.11 to 0.27. 5.3. Space heating performance The solar refrigeration system located at the tropical areas allows the refrigeration applications to be undertaken all the year around. However, in high latitudes with four obvious seasons of the year, we need to maximize the utilization of the system by conducting the cooling mode in summer and heating mode in winter. Thereby the solar energy utilization efficiency will be improved. In energy conversion process, the system heating mode was relatively simple compared to the cooling mode and the heat gained by PTC can be expressed with The energy conversion process of the system was relatively simple in the heating mode compared to the cooling mode, and the gained heat of the system can be expressed as Eq. (8). In the heating mode, the refrigeration unit (cooling chiller) acted as the heat exchanger. The output heat power from the cooling chiller can be expressed as Eq. (12). The system heating performance are tested with experiment and the results are shown in table 9 and the system characteristic parameter variation curves are given in Fig. 10 and Fig. 11. Where, Fig. 10 gives the variations of the running temperature of the absorption chiller, Where T7 and T8 are the inlet and outlet temperatures of the heat exchanger, respectively and the measurement positions of T5 to T8 are shown in Fig. 1. Fig. 11 gives the changes of indoor temperatures. Where, Tr1, Tr2 and Tr3 are the indoor temperatures and Ta (T15 in Fig. 1) is the corridor temperature. It shows that the indoor temperatures increased very quickly at the first 10 minutes of heat supply process. Table 9 gives the heat collected by PTC under different weather conditions. It shows that the average gained heat rate was about 0.4. In the practical applications, the total gained heat of the system can be estimated as Q h te , ins , s E i , tot (19) 5.4. Improvement analysis and discussion
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The experimental results revealed that the operating temperature and cooling temperature had a huge influence on the cooling performance of cooling system. Fig. 12 shows the changes of refrigeration coefficient at different running temperatures. The refrigeration quantity was increasing along with the increase of the hot water temperature. At the same time, the refrigeration coefficient was also substantially increasing with the increase of the hot water temperature. Where, T5 is the hot water temperature at inlet as shown in Fig. 1 and Table 1. In absorption refrigeration system, the function of cooling water is to exclude the absorption heat in the absorber and condensation heat in the condenser. Therefore, the intensity degree of absorption and condensation is related to temperature difference between inlet and outlet of cooling water in condenser and absorber. Fig. 13 shows the cooling performance influent by cooling water at different water temperatures. Where, T9 and T10 are the inlet and outlet temperatures of cooling water, respectively (see Fig. 1 and Table 1). Because the cooling pump was working in intermittent, the temperature of cooling water was displays in the shape of fluctuating with time. It can be also seen that when the cooling water temperatures reached to the maximum value, the temperature differences between inlet and outlet of cooling water was about 1.4ºC, while when the cooling water temperature down to the minimum value, the temperature differences between inlet and outlet of cooling water was about 3ºC. Therefore, when the inlet temperature of cooling water was 21ºC, the chiller refrigeration coefficient was about 0.68, when the inlet temperature of cooling water was 27.6ºC the refrigeration coefficient dropped to about 0.39. The refrigeration coefficient of the system is very close to that of the reference [21]. And the operation model was similar to that mentioned in reference [28]. In other words, in order to increase the refrigeration efficiency and refrigeration quantity, the temperature of cooling water should be as low as possible. In order to increase the comprehensive efficiency of the whole system (including refrigeration efficiency and heating efficiency), a certain operation conditions should be conducted as followings. Firstly, the area of PTC should be increased. Subsequently, temperature of the hot water and the quantity of heat storage will be improved. Thus, the cooling chiller can be started in a short period of time and operated at a capacity of 23 kW for a long running time. Secondly, the energy losses of each part of energy conversion materials should be reduced as far as possible, such as pipeline heat loss, hot water tank heat loss and so on. According to Table 6 the heat losses of system pipeline were relatively high; the main reason is that the length of pipelines among PTC, hot water tank and cooling chiller was too large. Therefore, to reduce the pipelines heat losses, the length of these pipelines should be reduced as possible. In order to tackle the heat losses problem, every part of the heat transfer pipelines should be shortened. According to the experimental measurements, the length of pipeline between PTC and hot water tank can be reduced to 50 m; the pipeline between hot water tank and cooling chiller can be shortened to 5 m. Thus, the heat losses between PTC and hot water tank will be approximately decreased from 3.90 - 6.2 kW to 1.40 - 2.30 kW. In addition, the heat losses between hot water tank and the cooling chiller will be also reduced from 0.90 - 1.40 kW to 0.25 - 0.40 kW. Subsequently, the gained heat of the whole system will be increased by 3.15 - 4.90 kW and the average COP of the whole system will be more than 0.3. Because the system was located in a windy area, so in order to improve system performance, system design optimization
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work must be carried out. One of the ways is that both of hot water storage tank and the cooling chiller should be installed in a small room so as to minimize the convection heat loss of them. Thirdly, to increase the chiller refrigeration quantity Qr and the COP of the whole cycle, the temperature of cooling water should be as low as possible. So the way of continuous cooling should be used. According to Fig. 13, when continuous cooling is adopted, the refrigeration efficiency immediately raised to 0.7. On the contrary, in intermittent cooling model, the average refrigeration efficiency was only about 0.55. 6. Conclusion The cooling performance of the single-effect LiBr–H2O absorption chiller driven by PTC has been investigated. The results revealed that the chiller’s average refrigeration coefficient ηr,av was between 0.18 and 0.60, and the average COP of the whole refrigeration cycle COPs,av was from 0.11 to 0.27 under different weather conditions. Moreover, the refrigeration quantity of the refrigeration chiller has been increased with the increasing of heat temperature. Meanwhile the refrigeration coefficient also substantially increased with the rising of hot water temperature. The cooling water temperature could largely influence on the performance of the refrigeration chiller. The values of ηte,ins,s and Pte,ins,s were relatively low, due to the pipeline was too long and its heat loss was too high. In addition, the values of ηr,av and COPs,av were also relatively low because the hot water temperature was too low. Therefore, in order to improve the cooling performance of the existing system the length of heat transfer pipelines should be shorten to reduce the heat losses. Furthermore, area of PTC and the operating temperature of the aperture should be increased and the water temperature for cooling system should be as low as possible. Acknowledgement The present study was supported by National Natural Science Foundation, China (Grant No.: U1137605). References [1] Otanicar T, Taylor RA, Phelan PE. Prospects for solar cooling – An economic and environmental assessment. Solar Energy. 2012;86:1287-99. [2] Siddiqui MU, Said SAM. A review of solar powered absorption systems. Renewable and Sustainable Energy Reviews. 2015;42:93-115. [3] Hassan HZ, Mohamad AA. A review on solar cold production through absorption technology. Renewable and Sustainable Energy Reviews. 2012;16:5331-48. [4] Kim DS, Infante Ferreira CA. Solar refrigeration options- a state-of-the-art review. International Journal of Refrigeration. 2008;31:3-15. [5] Iranmanesh A, Mehrabian MA. Dynamic simulation of a single-effect LiBr-H2O absorption refrigeration cycle considering the effects of thermal masses. Energy and Buildings. 2013;60:47-59. [6] Izquierdo M, González-Gil A, Palacios E. Solar-powered single-and double-effect directly air-cooled LiBr-H2O absorption prototype built as a single unit. Applied Energy. 2014;130:7-19. [7] Lizarte R, Izquierdo M, Marcos JD, Palacios E. Experimental comparison of two solar-driven air-cooled LiBr-H2O absorption chillers: Indirect versus direct air-cooled system. Energy and Buildings. 2013;62:323-34. [8] Agyenim F, Knight I, Rhodes M. Design and experimental testing of the performance of an outdoor LiBr-H2O solar thermal absorption cooling system with a cold store. Solar Energy. 2010;84:735-44. [9] Lamine CM, Said Z. Energy Analysis of Single Effect Absorption Chiller (LiBr-H2O) in an Industrial Manufacturing of Detergent. Energy Procedia. 2014;50:105-12.
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[10] González-Gil A, Izquierdo M, Marcos JD, Palacios E. Experimental evaluation of a direct air-cooled lithium bromide–water absorption prototype for solar air conditioning. Applied Thermal Engineering. 2011;31:3358-68. [11] Li Z, Ye X, Liu J. Performance analysis of solar air cooled double effect LiBr-H2O absorption cooling system in subtropical city. Energy Conversion and Management. 2014;85:302-12. [12] Iranmanesh A, Mehrabian MA. Optimization of a lithium bromide–water solar absorption cooling system with evacuated tube collectors using the genetic algorithm. Energy and Buildings. 2014;85:427-35. [13] López-Villada J, Ayou DS, Bruno JC, Coronas A. Modelling, simulation and analysis of solar absorption power-cooling systems. International Journal of Refrigeration. 2014;39:125-36. [14] Domínguez-Inzunza LA, Hernández-Magallanes JA, Sandoval-Reyes M, Rivera W. Comparison of the performance of singleeffect, half-effect, double-effect in series and inverse and triple-effect absorption cooling systems operating with the NH3– LiNO3 mixture. Applied Thermal Engineering. 2014;66:612-20. [15] Li Z, Ye X, Liu J. Optimal temperature of collector for solar double effect LiBr/H2O absorption cooling system in subtropical city based on a year round meteorological data. Applied Thermal Engineering. 2014;69:19-28. [16] Avanessian T, Ameri M. Energy, exergy, and economic analysis of single and double effect LiBr-H2O absorption chillers. Energy and Buildings. 2014;73:26-36. [17] Said SAM, El-Shaarawi MAI, Siddiqui MU. Analysis of a solar powered absorption system. Energy Conversion and Management. 2015;97:243-52. [18] Fong KF, Chow TT, Lee CK, Lin Z, Chan LS. Solar hybrid cooling system for high-tech offices in subtropical climate-Radiant cooling by absorption refrigeration and desiccant dehumidification. Energy Conversion and Management. 2011;52:2883-94. [19] Ketjoy N, yongphayoon R, Mansiri K. Performance Evaluation of 35 kW LiBr-H2O Solar Absorption Cooling System in Thailand. Energy Procedia. 2013;34:198-210. [20] Lu ZS, Wang RZ, Xia ZZ, Lu XR, Yang CB, Ma YC, et al. Study of a novel solar adsorption cooling system and a solar absorption cooling system with new CPC collectors. Renewable Energy. 2013;50:299-306. [21] Ali AHH, Noeres P, Pollerberg C. Performance assessment of an integrated free cooling and solar powered single-effect lithium bromide-water absorption chiller. Solar Energy. 2008;82:1021-30. [22] Hang Y, Qu M, Winston R, Jiang L, Widyolar B, Poiry H. Experimental based energy performance analysis and life cycle assessment for solar absorption cooling system at University of Californian, Merced. Energy and Buildings. 2014;82:746-57. [23] Yin YL, Zhai XQ, Wang RZ. Experimental investigation and performance analysis of a mini-type solar absorption cooling system. Applied Thermal Engineering. 2013;59:267-77. [24] Lazzarin RM, Romagnoni P, Casasola L. Two years of operation of a large solar cooling plant. International Journal of Refrigeration. 1993;16:185-90. [25] Xu SM, Huang XD, Du R. An investigation of the solar powered absorption refrigeration system with advanced energy storage technology. Solar Energy. 2011;85:1794-804. [26] Venegas M, Rodríguez-Hidalgo MC, Salgado R, Lecuona A, Rodríguez P, Gutiérrez G. Experimental diagnosis of the influence of operational variables on the performance of a solar absorption cooling system. Applied Energy. 2011;88:1447-54. [27] Al-Ugla AA, El-Shaarawi MAI, Said SAM. Alternative designs for a 24-hours operating solar-powered LiBr-water absorption air-conditioning technology. International Journal of Refrigeration. 2015;53:90-100. [28] Prasartkaew B. Performance Test of a Small Size LiBr-H2O Absorption Chiller. Energy Procedia. 2014;56:487-97. [29] Calise F, Palombo A, Vanoli L. Maximization of primary energy savings of solar heating and cooling systems by transient simulations and computer design of experiments. Applied Energy. 2010;87:524-40. [30] Noro M, Lazzarin RM. Solar cooling between thermal and photovoltaic: An energy and economic comparative study in the Mediterranean conditions. Energy. 2014;73:453-64.
Page 14 of 24
Wind velocity sensor
Solar beam radiation meter
Blower coil
Data acquisition system
Blower coil
Blower coil
Meeting room (102 m2) T15
T14
T13
Supplement water tank
Make up water Supplement water tank
T12 T4
T3
Hot water storage tank
T8
T6
TX-23
T11
PTC2 T5
T2
T1
Chilled water tank
Make up water
Pollution Discharge vavle
Hot water pump
LiBr absorption cooling chiller
T7 Pollution Chilled water pump discharge vavle T10 Cooling tower
T9 Cooling water pump
PTC1
Pump
Pollution discharge vavle
Pollution discharge vavle
Valve
Fig. 1 – Scheme diagram of solar powered absorption cooling system.
Page 15 of 24
Fig. 2 –Main parts of the solar absorption cooling system.
Sun
Qp,loss,1
Ei Qtc,ins PTC array
Pipeline
Hot water Storage tank
pipeline
The end of air-condition
Qa,room
Pipeline
Qp,a,3
Absorption cooling chiller
Qt,loss
Qp,loss,2
Cooling tower
Qe
Qc
Fig. 3 – Heat flows diagram of the solar absorption cooling system.
Page 16 of 24
0.8
35
0.9
0.7
30
0.8
0.4 15
0.3 te,ins
0.2
10
Ptc,ins Ib
0.1 0.0 12.5 12:30
13.0 13:00
13.5 13:30
14.0 14:00
14.5 14:30
0.5 0.4 0.3
5
15.0 15:00
0.6
-2
20
Ptc,ins (kW)
te,ins
0.7
25
0.5
Ib (kW m )
0.6
0.2
0 15.5 15:30
0.1
Time (h) Time (hh:mm) Time (hh:mm)
Fig. 4 – Test results of instantaneous efficiency and instantaneous power of PTC array.
0.7
35
0.9
0.6
30
0.8
0.5
25
0.4
20
0.3
15
0.2
10
te,ins,s
-2
0.5
Ib (kW m )
0.6
0.4 0.3 0.2
Ptc,ins,s Ib
0.1
0.7
Ptc,ins,s (kW)
te,ins,s
Fig. 4 – Test results of instantaneous efficiency and instantaneous power of PTC array.
5
0.0 0 12.5 12:30 13.0 13:00 13.5 13:30 14.0 14:00 14.5 14:30 15.0 15:00 15.5 15:30
0.1 0.0
Time Time (h)(hh:mm) Fig. 5 – Thermal efficiency and power of heat collection system.
Fig. 5 – Thermal efficiency and power of heat collection system. 1.0
5
0.8
4
0.4
3
0.3
2
0.2 1
0.1 0.0
8:00 8
10:00 10
12:00 12
14:00 14
Time (hh:mm) Time (h)
(a) 2014-05-11
16:00 16
0
18:00 18
2
0.5
8 Ib
vw
7 6 5
0.6
4 0.4
3 2
0.2 0.0
8:00 8
vw (m/s)
vw
6
2
Ib (kW/m )
0.6
Ib
Ib (kW/m )
0.7
vw (m/s)
0.8
1 10:00 10
12:00 12
14:00 14
16:00 16
0
18:00 18
Time (hh:mm) Time (h)
(b) 2014-05-14
Page 17 of 24
Fig. 6 – Solar beam radiation and wind velocity vs. time. T5 T7 T9
70 60
T6 T8 T10
70 60 50
T ( C)
40
o
o
T( C)
50
30
30
20
20
10
12:54
13:54
14:54
15:54
10
16:54
T5 T7 T9
40
T6 T9 T10
12:19 13:19 14:19 15:19 16:19 17:19 18:19
Time (hh:mm)
Time (hh:mm)
(a) 2014-05-11
(b) 2014-05-14
0.6
16 r
r
14
Pr
0.5
1.0
6
0.2
0.6
r
8
Pr (kW)
10
0.3
4 0.1 0.0
0.2
2 13 13:00
14 14:00
13:00
14:00
15 15:00
16 16:00
17 17:00
0 18 18:00
17:00
18:00
0.4
0.0 12:00 12 13:00 13 14:00 14 15:00 15 16:00 16 17:00 17 18:00 18 19:00 19
Time (h) Time (hh:mm) 15:00
Time (h) Time (hh:mm)
16:00
(a) 2014-05-11
Time Time (hh:mm) (hh:mm)
(b) 2014-05-14
Fig. 8 – The refrigeration coefficient and refrigeration power of the absorption cooling chiller. 28
24 23
26
22
24
T ( C)
20
o
T ( C)
21 o
r
0.4
Pr
0.8
12
22 20 18 16 14 12 10 8 6 4 2 0
Pr (kW)
Fig. 7 – The inlet and outlet temperatures of heating, chilled and cooling of chiller vs. time.
19 18
Tr1 Tr2 Ta
17 16
12:54
13:54
14:54
15:54
16:54
22 Tr1 Tr2 Ta
20 18
12:19 13:19 14:19 15:19 16:19 17:19 18:19
Time (hh:mm)
Time (hh:mm)
(a) 2014-05-11
(b) 2014-05-14
Page 18 of 24
Fig. 9 – The variation of indoor temperatures during refrigerating. 50 45
35
o
T ( C)
40
30
T5 T6 T7 T8
25 20
15:48 15.8 16:00 16.0 16:12 16.2 16:24 16.4 16:36 16.6 16:48 16.8 17:00 17.0 17:12 17.2 17:24 17.4
Time (h) Time (hh:mm) Fig. 10 – The running temperatures of absorption chiller when it was under the heating mode (Mar. 29). Fig. 10 – The running temperatures of absorption chiller when it was under the heating mode (Mar. 29). 28 26
o
T ( C)
24 22 20
Tr1 Tr3
Tr2 Ta
18 15:48 15.8 16:00 16.0 16:12 16.2 16:24 16.4 16:36 16.6 16:48 16.8 17:00 17.0 17:12 17.2 17:24 17.4
(hh:mm) Time (h) Time (hh:mm) Fig. 11 –running The variations of room temperatures during theitheating mode 29).mode (Mar. 29). Fig. 10 – The temperatures of absorption chiller when was under the(Mar. heating
Fig. 11 – The variations of room temperatures during the heating mode (Mar. 29). 22 20
r
Pr
0.8
18
0.7
16
0.6
14 12
0.5
r
10
0.4
8
0.3
6
0.2
4 2
0.1 0.0 46
Pr (kW)
0.9
0 48
50
52
54
56
58
60
62
64
66
68
70
72
74
o
T5 ( C)
Page 19 of 24
Fig. 12 – Instantaneous refrigeration coefficient vs. running temperatures. Fig. 12 – Refrigeration coefficient vs. running temperatures.
31
0.75 T9,
30
T10,
r
0.70
29
0.65
28
0.60
26
0.55
25
0.50
24 23
0.45
22
0.40
21 20 13:36 13.6
r
o
T ( C)
27
13:48 13.8
14:00 14.0
14:12 14.2
14:24 14.4
14:36 14.6
14:48 14.8
0.35
15:00 15.0
Time (h) Time (hh:mm)
Fig. 13 – Cooling performance influent by cooling water temperature.
Fig. 13 – Cooling performance influent by cooling water temperature. Table 1 – Instructions for the temperature probes in Fig. 1. Temperature Measurement explain probe T1
Inlet temperature of PTC1
T2
Outlet temperature of PTC1
T3
Inlet temperature of PTC2
T4
Outlet temperature of PTC2
T5
Hot water inlet temperature of chiller
T6
Hot water outlet temperature of chiller
T7
Chilled water inlet temperature of chiller
T8
Chilled water outlet temperature of chiller
T9
Cooling water inlet temperature of chiller
T10
Cooling water outlet temperature of chiller
T11
Outlet temperature of hot water tank
T12
Inlet temperature of hot water tank
T13, T14
The indoor temperature of meeting room
T15
The outdoor temperature of meeting room
Page 20 of 24
Table 2- Main measurements and instruments accuracy of the monitoring system. Instrument
Model
Range
Accuracy
Application scope
Maximum relative error
Maximum absolute error
Uncertainty (B class)
Pyranometer
Kipp & Zonen CMP-6
0-2000 2 (W/m )
±5%
0-1000 2 (W/m )
±10%
±100W/m2
57.7348 2 W/m
Thermocouples
T
-200 to 350 (℃)
±0.4%
0-150 (℃)
±0.93%
±1.4℃
0.8083℃
Wind speed transducer
EC-9S
070(m/s)
±0.4%
0-10 (m/s)
±2.8%
±0.28m/s
0.1617 m/s
Electromagnetic flow meter
KROHNE OPTIFLUX 5300
DN 25; 012 (m/s)
±0.15%
0-5 (m/s)
±0.36%
±0.018m/s
0.0104 m/s
Pressure transducer
YOKOGAWA EJA430E
0.14-16 (MPa)
±0.055%
0-2 (MPa)
±0.44%
±0.0088MPa
0.0051 MPa
Table 3 – PTC array parameters. Parameter
Value
Aperture area
56 m
Aperture width
2.5 m
Orientation
North-South (ψ = 0°)
Length of PTC
26 m
Focal distance of parabolic trough
1.1 m
The width of focal spot
5 cm
PTC efficiency values at different temperature
77.5%(30°C), 68.4%(50°C), 57.3%(70°C), 38.4%(90°C)
Inner diameter of metal pipe of receiver
4 cm
Outer diameter of glass pipe of receiver
11 cm
2
Page 21 of 24
Table 4 – Parameters of absorption chiller. Parameter
Value
Model
TX-23
Working pair
Lithium Bromide-Water
Ambient temperature
28°C -36.0°C
hot water temperature at chiller inlet
50°C -90.0°C
hot water temperature at chiller outlet
49°C -80.0°C
Outlet temperature of chilled water
10.0°C-14.0°C
Inlet temperature of chilled water
15.0°C-22.0°C
Cooling water temperature at chiller inlet
20.7°C-27.8°C
Cooling water temperature at chiller outlet
23.5°C-29.3°C
Power dissipation
2.3 kW
Refrigeration capacity
23 kW
Chilled water rate
4.0 m3/h
Hot water rate
5.7 m3/h
Table 5 – Parameters of cooling tower, tank and pumps. Parameter
Value
Model
BLT-10
Wind rate
10.5 m /h
Cooling water rate
10 m /h
Dynamo power
0.75 kW
Capacity of tank
1m3
Insulation thickness of tank
3 cm
Pump of PTC
0.33 kW
Hot water pump
0.78 kW
Chilled water pump
0.78 kW
Cooling water pump
0.98 kW
Comment [A1]: Author: There are two different Table 5 captions were provided in the manuscript, and this has been retained. Please check and confirm it is correct.
3
3
Page 22 of 24
Table 6 – Parameters of blower coil. Parameter
Value
Model
EKCW800KT
Wind rate
1360 m /h
Cooling power
7200 W
Input power
130 W
3
Table 7 – Parameter of pipes and the energy losses of the pipelines. Pipeline type
Pipeline between PTC and hot water storage tank Pipeline between hot water storage tank and chiller Pipeline between chiller and blower coil
Inner tube diameter
Outer tube diameter
Insulation thickness
Pipe length
Total energy losses (kW)
4 cm
12.5 cm
4 cm
78 m
3.90–6.20
4cm
12.5 cm
4 cm
18 m
0.90–1.40
4cm
10 cm
3 cm
55 m
0.27–0.55
Note
Comment [A2]: Author: There are two different Table 7 captions were provided in the manuscript, and this has been retained. Please check and confirm it is correct.
60–90 ºC (heat loss) 60–90 ºC (heat loss) 7–16 ºC (cooling loss)
Table 8 – The average coefficient of performance for the refrigeration system. Date (2014)
Weather condition
Ei,tot (MJ)
Qr (MJ)
ηr,av
COPs,av
Apr. 4
Cloudy and windy
448
54
0.18
0.12
May 10
Cloudy and windy
643
84
0.21
0.13
May 11
Cloudy and windy
549
58
0.17
0.11
May 13
Sunny and windy
1300
220
0.51
0.17
May 14
Sunny and windy
1200
170
0.45
0.14
May 15
Sunny and windy
1122
234
0.57
0.21
May 16
Sunny and windy
1120
267
0.58
0.23
May 17
Sunny and windy
1280
284
0.51
0.22
May 18
Sunny and windy
1262
313
0.55
0.25
Page 23 of 24
May 19
Sunny and windy
877
288
0.59
0.23
May 20
Cloudy and windy
690
158
0.51
0.21
May 21
Sunny and windy
967
270
0.57
0.27
May 22
Sunny and windy
1110
262
0.47
0.24
May 24
Sunny and windy
1084
234
0.50
0.22
Table 9
– The gained heat of system in the heating mode.
Date (2014)
Weather condition
Ei,tot (MJ)
Qh (MJ)
Average gained heat rate
Mar. 2
Sunny and windy
1032
412
0.399
Mar. 29
Cloudy and windy
423
164
0.387
May 12
Cloudy and windy
197
81
0.411
Page 24 of 24