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Investigating the energy performance of an air treatment incorporated cooling system for hot and humid climate
Accepted Manuscript Title: Investigating the energy performance of an air treatment incorporated cooling system for hot and humid climate Authors: X. Cui, B. Mohan, M.R. Islam, K.J. Chua PII: DOI: Reference:
S0378-7788(16)30945-8 http://dx.doi.org/doi:10.1016/j.enbuild.2017.06.059 ENB 7722
To appear in:
ENB
Received date: Revised date: Accepted date:
24-9-2016 14-2-2017 21-6-2017
Please cite this article as: X.Cui, B.Mohan, M.R.Islam, K.J.Chua, Investigating the energy performance of an air treatment incorporated cooling system for hot and humid climate, Energy and Buildingshttp://dx.doi.org/10.1016/j.enbuild.2017.06.059 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.
Investigating the energy performance of an air treatment incorporated cooling system for hot and humid climate X. Cui1, B. Mohan 1, M.R. Islam1,2, K.J. Chua1,2 * 1
Department of Mechanical Engineering, National University of Singapore, 9 Engineering Drive 1, Singapore 117576, Singapore 2 Engineering Science Programme, National University of Singapore, 9 Engineering Drive 1, Singapore 117575, Singapore *Tel. + 65 6516 2558; fax: + 65 6779 1459. E-mail address: [email protected] (K.J. Chua)
Highlights
An air treatment system has been introduced for enhanced indoor air quality.
A lower outdoor air flow rate is employed due to the air-purification process.
A mathematical model for the cooling coil unit has been developed.
Experimental systems have been constructed to study the real-time performance.
The system could achieve marked energy savings in hot and humid climate.
Abstract: An air treatment system incorporating an air-purification unit has been proposed to reduce the energy consumption of air-conditioning system and improve indoor air quality. This system employs a lower outdoor-air fraction to realize the acceptable indoor air quality and thermal comfort resulting in marked energy savings due to the reduced cooling load for the outdoor air. The performance of the primary equipment of the air treatment system has been individually investigated. Experimental systems have been constructed to study the air-purification process and to evaluate the chiller’s performance. Experimental results have demonstrated that the proposed air-purification unit is able to remove indoor air pollutants such as volatile organic compound. A computational model has concurrently been developed to determine the performance of the air handling cooling coil which is
impacted by the outdoor-air fraction and the chilled water temperature. A higher chilled water supply temperature is adopted by reducing the outdoor-air fraction without compromising the supply air cooling capacity. Having a reduced outdoor air cooling load and a better chiller’s energy efficiency operating with higher chilled water temperatures, the achievable energy savings of the airconditioning system has been demonstrated to be up to 36% under a tropical hot and humid climate.
Keywords: Air-conditioning; Heat transfer; Mathematical model; Energy consumption.
Nomenclature A cp D dp Ga h hm i H k Lef m P Pr Q Re s Sl St T u V Wf μ
ν ρ 𝜖 𝜂𝑠
area [m2] specific heat [kJ/(kg °C)] tube diameter [m] particle mean diameter air flow rate per square meter [kg/(s· m2)] convection heat transfer coefficient [kW/(m2 °C)] mass transfer coefficient [kg/(m2 s)] enthalpy of air [kJ/kg] length of the adsorption column [m] thermal conductivity [W/(m·°C)] Lewis factor mass flow rate [kg/s] pressure [kPa] Prandtl number heat transfer rate [kW] Reynolds number fin spacing [m] tube spacing in air flow direction [m] tube spacing normal to air flow [m] temperature [°C] velocity [m/s] volume flow rate of air [m3/s] fan power [kW] dynamic viscosity coefficient [Pa·s] kinematic viscosity [m2/s] density [kg/m3] void fraction surface effectiveness 2
ω
humidity ratio [kg/kg]
Subscript a air dew dew-point temperature of air i inner surface in inlet of cooling coil m mean value o outer surface out outlet of cooling coil s surface w water
1. Introduction
Outdoor-air ventilation is a common method to maintain a healthy indoor air. Increasing the ventilation rate of the air-conditioning system through greater outdoor air-intake has a positive impact on reducing sick-building syndrome (SBS) symptoms [1,2]. However, increasing the outdoor air fraction more often than not brings about greater energy consumption attributed to a higher cooling load for a higher level of fresh outdoor air, especially in hot and humid climates [3]. Therefore, it is an ongoing research challenge to find alternative cooling and ventilation strategies to reduce energy consumption and improve indoor air quality. To achieve energy efficient air conditioning systems, researchers have conducted extensive works in terms of designing and engineering novel systems [4–7]. Lin et al. [8] compared year-round energy consumptions of three ventilation methods for typical indoor spaces in Hong Kong. It was observed that the stratum ventilation provided greater energy savings when compared with mixing ventilation and displacement ventilation. Fong et al. [9] further illustrated that the rear-middle-level-
3
exhausted stratum ventilation provided a satisfied thermal comfort condition with least energy consumption. Chua et al. [10] conducted a comparison study on the dehumidification performances of three temperature control strategies, namely, chilled water flow control, bypass air control and the variable air volume control. Theoretical analysis on the cooling coil was developed to investigate the part-load performance. Niu et al. [11] proposed an air-conditioning system that combined chilledceiling with desiccant cooling for hot and humid climates. Their results highlighted that the combined system was able to save up to 44% of primary energy consumption compared with a conventional constant volume all-air system. Another low-energy air-conditioning strategy was reported to combine the employment of chilled-ceiling with the use of microencapsulated phase change material slurry storage and evaporative cooling technology to promote energy savings [12]. To curb energy consumption, a number of researchers have proposed combining the conventional airconditioning system with energy-efficient devices[13–15]. Younis et al. [16] investigated the performance of a displacement ventilation system combined with an evaporative cooled ceiling. Kim et al. [17] employed liquid desiccant dehumidification devices and evaporative coolers to enhance energy efficiency. Lee et al. [18] conducted a theoretical study on the decoupling sensible and latent cooling for desirable indoor air conditions. The results indicated that the annual energy consumption could be reduced by 54% over conventional system. In general, the energy-efficient systems reviewed relied on fresh outdoor air to maintain a desired indoor air quality. However, it may be undesirable to sustain a constant level of outdoor air intake if the outdoor air quality is poor. When such a situation occurs, the use of an air purification process is imperative to continue to sustain good indoor air quality without penalizing the energy 4
efficiency of the air-conditioning system. In modern buildings, volatile organic compound (VOC) is known as one of the primary indoor airborne pollutants [19]. It has been reported that some of VOCs may cause potential adverse health issues [20,21]. Several studies have proposed possible strategies for removing VOCs. Mo et al. [22] developed a general model for analyzing VOC removal performance of photocatalytic oxidation reactors. Cheng et al. [23] compared several building materials in terms of their VOC emissions. Building materials with mineral content obtained generated the least byproducts after the reaction with ozone. Waring and Wells [24] studied the impact of using ozone, hydroxyl radical, and the nitrate radical on indoor residential VOC conversion. They indicated that the total VOC conversion was dominated by ozonolysis and the hydroxyl radical reactions. Sidheswaran et al. [25] investigated the air cleaning performance of using activated carbon fiber (ACF) filters in air-conditioning systems. The ACF media was able to adsorb VOCs from indoor air. During the regeneration process, VOCs can be desorbed from the ACF media so that it can be used in the next cycle of air cleaning. Thus far, existing works have yet to adequately evaluate the energy performance of an airconditioning system equipped with an air-purification unit for minimizing the outdoor air fraction in hot and humid climate. In addition, in recent years, Singapore has been suffering from severe smoke haze regularly due to the numerous forest fires in Southeast Asian. The haze effect is particularly prominent during the Southwest Monsoon Season [26]. This situation further spurs us to developing innovative methods to improve energy efficiency of existing air-conditioning systems while reducing the polluted outdoor air intake and maintaining an acceptable indoor air quality. To address these issues, the present work proposes an air treatment system with an air5
purification unit for reduced cooling load and improved air quality in hot and humid climate like Singapore. The key objectives of this work are to (1) investigate the performance of individual primary equipment of an innovative air treatment system that employs ozone, and (2) evaluate the energy performance of the air-conditioning system with regulated outdoor air flow rate. We will first introduce the air-treatment system design. Thereafter, a mathematical modelling of the air conditioning system’s chilled water cooling coil will be developed, followed by the description of experimental setup that incorporates key components. Based on the performance of the air handling unit and the incorporated ozone air purifier, we further illustrate varying levels of reducing energy consumption of the air-conditioning system.
2. Description of the air treatment system
Fig. 1 presents the schematic of the proposed ozone-based air treatment system to improve air quality and to reduce energy consumption. The return air is first filtered and purified in an ozonebased oxidation device. The ozone is produced from the generator using oxygen which is generated by the combined electrolysis and photocatalysis processes. The oxygen can be used to enrich the supply air as well. Thereafter, the CO2 scrubber, filled with activated carbon filter, reduces the CO2 concentration and further purifies the supply air. Finally, the air is cooled and dehumidified in air handling unit using chilled water. Due to the air-purification and oxygen enrichment processes, this system is able to employ a lower outdoor-air fraction resulting in a significant saving on energy consumption of the air conditioning system. 6
3. Mathematical formulation
3.1 Cooling coil of air handling unit
The chilled water coil of air handling unit is an important part of the overall cooling system to maintain the comfortable indoor thermal condition. The supplied air is cooled and dehumidified by passing it through the cooling coil. To precisely determine the performance of the cooling coil, a mathematical model is developed and judiciously described in this section. The calculation is carried out using the “row-by-row method” [27,28]. The computational domain is discretized into N segments for each row of the cooling coil. Fig. 2 illustrates the computational element for one segment in which governing equations are established. For the air side, the total heat transfer can be evaluated by the enthalpy change: ∆𝑄(𝑖,𝑗) = 𝑚𝑎 (𝑖𝑎(𝑖,𝑗) − 𝑖𝑎(𝑖+1,𝑗) )
(1)
In addition, the total heat transfer is a sum of the sensible heat transfer and the latent heat transfer: ∆𝑄(𝑖,𝑗) = ∆𝑄𝑠(𝑖,𝑗) + ∆𝑄𝑙(𝑖,𝑗) = ℎ𝑜 𝛥𝐴𝑜 𝜂𝑠 (𝑇𝑎,𝑚(𝑖,𝑗) − 𝑇𝑠,𝑚(𝑖,𝑗) ) + ℎ𝑓𝑔 ℎ𝑚 𝛥𝐴𝑜 𝜂𝑠 (𝜔𝑎,𝑚(𝑖,𝑗) − 𝜔𝑠,𝑚(𝑖,𝑗) )
(2)
where 𝛥𝐴𝑜 is the outer surface area of cooling coil, and 𝜂𝑠 is the surface effectiveness. The external heat transfer coefficient can be obtained based on the Colburn j-factor analogy [29]: ℎ𝑜 = 𝑗𝐺𝑎 𝑐𝑝𝑎 𝑃𝑟 −2/3
7
(3)
where 𝑆
𝑗 = 0.14𝑅𝑒𝑎−0.328 ( 𝑡)
−0.502
𝑆𝑙
𝑠
( )
0.0312
(4)
𝐷𝑜
The Lewis factor is approximately unity for air/water mixtures 𝐿𝑒𝑓 =
ℎ𝑜
≈1
(5)
ℎ𝑜 Δ𝐴𝑜 (𝑖𝑎,𝑚(𝑖,𝑗) − 𝑖𝑠,𝑚(𝑖,𝑗) )
(6)
ℎ𝑚 ∙𝑐𝑝𝑎
Therefore, Eq. (2) can be rewritten as: ∆𝑄(𝑖,𝑗) =
𝜂𝑠 𝑐𝑝𝑎
For the chilled water, the total heat transfer can be expressed similarly as: ∆𝑄(𝑖,𝑗) = 𝑚𝑤 𝑐𝑝𝑤 (𝑇𝑤(𝑖,𝑗+1) − 𝑇𝑤(𝑖,𝑗) )
(7)
∆𝑄(𝑖,𝑗) = ℎ𝑖 ∆𝐴𝑖 (𝑇𝑠,𝑚(𝑖,𝑗) − 𝑇𝑤,𝑚(𝑖,𝑗) )
(8)
The inner heat transfer coefficient for the chilled water side can be determined by the correlation described by Petukhov [30]: 2
𝑓
𝑁𝑢𝑖 =
(8)(𝑅𝑒𝑤 −1000)𝑃𝑟𝑤
2 𝑓 0.5 3 −1) 1+12.7(8) (𝑃𝑟𝑤
[1 +
𝐷 3 ( 𝑖) ] 𝐿
(9)
where 𝑓 = (0.79 𝑙𝑛 𝑅𝑒𝑤 − 1.64)−2
(10)
𝑖𝑎,𝑚(𝑖,𝑗) , 𝑖𝑠,𝑚(𝑖,𝑗) 𝑇𝑎,𝑚(𝑖,𝑗) , 𝑇𝑠,𝑚(𝑖,𝑗) , and 𝑇𝑤,𝑚(𝑖,𝑗) are the mean value between adjacent grid point. For example, 𝑖𝑎,𝑚(𝑖,𝑗) can be expressed as 𝑖𝑎,𝑚(𝑖,𝑗) =
(𝑖𝑎(𝑖,𝑗) +𝑖𝑎(𝑖+1,𝑗) ) 2
(11)
Eliminating 𝑖𝑎(𝑖+1,𝑗) by substituting 𝑖𝑎,𝑚(𝑖,𝑗) into Eq. (1), the energy equations become: ∆𝑄(𝑖,𝑗) = 2𝑚𝑎 (𝑖𝑎(𝑖,𝑗) − 𝑖𝑎,𝑚(𝑖,𝑗) ) 8
(12)
Similarly, eliminating 𝑇𝑤(𝑖+1,𝑗) by substituting 𝑇𝑤,𝑚(𝑖,𝑗) into Eq. (7): ∆𝑄(𝑖,𝑗) = 2𝑚𝑤 𝑐𝑝𝑤 (𝑇𝑤,𝑚(𝑖,𝑗) − 𝑇𝑤(𝑖,𝑗) )
(13)
Eliminating 𝑖𝑎,𝑚(𝑖,𝑗) between Eq. (6) and Eq. (12) yields ∆𝑄(𝑖,𝑗) =
𝜂𝑠 ℎ ∆𝐴𝑜 𝑐𝑝𝑚 𝑜
∙ (𝑖𝑎(𝑖,𝑗) − 𝑖𝑠,𝑚(𝑖,𝑗) )
1+∆𝑁𝑇𝑈𝑜 /2
(14)
where ∆𝑁𝑇𝑈𝑜 =
𝜂𝑠 ℎ𝑜 ∆𝐴𝑜 𝑚𝑎 𝑐𝑝𝑚
(15)
Again, eliminating 𝑇𝑤,𝑚(𝑖,𝑗) between Eq. (8) and Eq. (13) yields ∆𝑄(𝑖,𝑗) =
ℎ𝑖 ∆𝐴𝑖 1+∆𝑁𝑇𝑈𝑖 /2
∙ (𝑇𝑠,𝑚(𝑖,𝑗) − 𝑇𝑤(𝑖,𝑗) )
(16)
where ∆𝑁𝑇𝑈𝑖 =
ℎ𝑖 ∆𝐴𝑖 𝑚𝑤 𝑐𝑝𝑤
(17)
The air enthalpy on the coil surface (𝑖𝑠,𝑚(𝑖,𝑗) ) and the coil surface temperature (𝑇𝑠,𝑚(𝑖,𝑗) ) are essential parameters for calculating the ultimate temperature and humidity ratio of air. The coil surface condition depends largely on its surface temperature. When the coil surface temperature is higher than the dew point temperature of the air, the surface of the cooling coil remains dry so that the relationship between 𝑖𝑠,𝑚(𝑖,𝑗) and 𝑇𝑠,𝑚(𝑖,𝑗) is given as 𝑖𝑠,𝑚(𝑖,𝑗) = 𝑖𝑎(𝑖,𝑗) + 𝑐𝑝𝑎 (𝑇𝑠,𝑚(𝑖,𝑗) − 𝑇𝑎(𝑖,𝑗) ), 𝑇𝑠,𝑚(𝑖,𝑗) > 𝑇𝑑𝑒𝑤(𝑖,𝑗)
(18)
When the coil surface temperature is lower than the dew point temperature of the air, the coil surface is wet since the air is dehumidified by passing through the coil. In such case, 𝑖𝑠,𝑚(𝑖,𝑗) should
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be evaluated by the enthalpy of the saturated air at the coil outer surface temperature as follows: 2 𝑖𝑠,𝑚(𝑖,𝑗) = 10.76 + 1.4𝑇𝑠,𝑚(𝑖,𝑗) + 0.046𝑇𝑠,𝑚(𝑖,𝑗) , 𝑇𝑠,𝑚(𝑖,𝑗) ≤ 𝑇𝑑𝑒𝑤(𝑖,𝑗)
(19)
After the determination of 𝑇𝑠,𝑚(𝑖,𝑗) , the total heat transfer ∆𝑄(𝑖,𝑗) can be calculated using Eq. (16). Thereafter, the parameters for the following grid can be calculated as 𝑖𝑎(𝑖+1,𝑗) = 𝑖𝑎(𝑖,𝑗) − 𝑇𝑤(𝑖,𝑗+1) = 𝑇𝑤(𝑖,𝑗) −
∆𝑄(𝑖,𝑗)
(20)
𝑚𝑎 ∆𝑄(𝑖,𝑗)
(21)
𝑚𝑤 𝑐𝑝𝑤
In addition, by solving sensible heat transfer equations, we obtain 𝑇𝑎(𝑖+1,𝑗) =
∆𝑁𝑇𝑈𝑜 2 ∆𝑁𝑇𝑈 1+ 2 𝑜
1−
𝜔𝑎(𝑖+1,𝑗) =
∙ 𝑇𝑎(𝑖,𝑗) +
∆𝑁𝑇𝑈𝑜
∆𝑁𝑇𝑈𝑜 2
1+
∙ 𝑇𝑠,𝑚(𝑖,𝑗)
𝑖𝑎(𝑖+1,𝑗) −𝑐𝑝𝑎 ∙𝑇𝑎(𝑖+1,𝑗) [2501+1.89∙𝑇𝑎(𝑖+1,𝑗) ]∙1000
(22) (23)
3.2 Energy consumption of primary equipment
The electrical power consumed by the chiller is provided below as 𝑐ℎ𝑖𝑙𝑙𝑒𝑟 𝑝𝑜𝑤𝑒𝑟 =
𝑄𝑐 𝐶𝑂𝑃
(24)
where 𝑄𝑐 is the cooling load, and COP is the coefficient of performance of the chiller. The fan performance is calculated based on its operational air flow rate, pressure rise and efficiency. The total power is given as: 𝑊𝑓 =
𝑉∆𝑝 𝜂𝑏 𝜂𝑚
where 𝜂𝑏 is the blower efficiency, and 𝜂𝑚 is the motor efficiency. 10
(25)
The pressure drop through the activated carbon particles can be calculated as [31]: ∆𝑝 =
150𝜇𝐻 (1−𝜖)2 2 𝑑𝑝
𝜖3
𝑢+
1.75𝐿𝜌 (1−𝜖) 𝑑𝑝
𝜖3
𝑢2
(26)
where 𝑑𝑝 is the particle mean diameter, 𝜖 is the void fraction, and u is the air velocity.
4. Experimental setup
4.1 Air purification process
To independently evaluate the air purification performance of the ozone-based oxidation process, a separate experimental setup was fabricated as shown in Fig. 3. The experimental setup included an insulated air-tight chamber, an ozone generator, and sensors for measuring air parameters such as VOC concentration. The insulated chamber was made from stainless steel with size of 1m × 1m × 1m. A fan with variable speed drive was installed to mix the air inside the chamber. The ozone can be injected into the chamber through the ozone generator which was regulated by a mass flow controller. Sensors were inserted into the chamber to measure the concentration of various gases, such as VOC and Ozone. Table 1 indicates the specifications of measuring instruments. The following procedures are employed for testing the reaction between VOCs and ozone. Firstly, samples of building material or chemicals were placed inside the chamber for VOC emission. Secondly, the samples were then removed and the concentration of VOC was monitored until its concentration becomes relatively stable. Thirdly, a specific quantity of ozone was injected into the chamber. The ozone generator was turned off once a required ozone concentration was reached.
11
Consolidated data collected from sensors were monitored and recorded to observe the impact of ozone. Another key component of the air treatment system is the CO2 scrubbing unit. Fig. 4 depicts the experimental setup for CO2 scrubbing process. A chamber with a size of 750 mm × 750 mm × 750 mm was designed to simulate the indoor environment. The chamber was equipped with a small fan for mixing the air. The CO2 cylinder provided the CO2 gas into the chamber through a mass flow controller. The blower with variable speed drive was able to circulate the air at specific flow rate through the CO2 scrubber canister. In this study, the canister was filled with activated carbon which was used as the CO2 absorbing material. CO2 sensors monitored and recorded CO2 concentrations at various locations in order to investigate the effect of air flow velocity on CO2 scrubbing performance. The specifications of measuring instruments are shown in Table 1.
4.2 Chiller
Experiments were also carried out to analyze the performance of a chiller at different chilled water supply temperatures. Fig. 5 illustrates the schematic diagram of the experimental setup. It consisted of a water cooled scroll chiller of cooling capacity 15 kW, shell and tube type evaporator and condenser, two water mixing chambers, pumps, and a cooling tower. The inlet water temperature was regulated using two water mixing chambers for both the evaporator and the condenser. The chilled water supply temperature was varied from 7 to 13 °C. The cooling tower was installed to provide cooled water to be conditioned and circulated back to the mixing chambers.
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5. Results and discussion
In this section, we first validate the cooling coil mathematical model, followed by the experimental results on the air purification process and the performance of chiller. Based on the simulation results and experimental data, we further evaluate the thermal performance of the air handling unit and the energy consumption of the air treatment system. The framework of this study is illustrated in Fig. 6.
5.1 Model validation on cooling coil
The mathematical modeling was validated against experimental data acquired from a previous study [32]. Zhou and Braun have investigated the performance of both four-row and eight-row cooling coils [32]. Experimental study was conducted under a variety of operating conditions for inlet air flow velocity, inlet air temperature, inlet air humidity, water flow rate, and inlet water temperature. Table 2 indicates the boundary conditions for steady-state operations. The computational model was validated by comparing the simulated outlet conditions with experimental data. Fig. 7 demonstrates the capability of the computational model to predict the cooling coil performance. It portrayed a good agreement with experiments a discrepancy of less than ±10%.
5.2 Air purification performance
VOCs are usually emitted from building materials such as paint and glue. In this study, samples 13
of paint and glue were tested for VOC abatement using ozone in the experimental chamber. In addition, two pure chemicals, namely, Ethylbenzene and Xylene, were also treated in the chamber to investigate the effect of oxidation reaction. Fig. 8 presents the experimental results on VOCs treatment. The VOC for paint and glue are reduced by around 19% and 44%, respectively, over a time period of 24 hours. Comparing with paint sample, the glue sample demonstrates a faster reaction rate and consumes more ozone within the same time period. It is also observed that the reduction rate for Ethylbenzene and Xylene are 23% and 26%, respectively, over a time period of 30 hours. The experimental result shows the feasibility of using ozone to markedly reduce VOCs in an enclosed environment. Fig. 9 further shows the CO2 concentration reduction percentage in the experimental chamber. The initial CO2 concentration in the chamber was kept as around 1200 ppm. To evaluate the performance of canister in CO2 reduction, the air was circulated through the canister under varying velocities. It can be seen that a higher air flow velocity leads to improve CO2 reduction rate. For example, to obtain 15% reduction of CO2, it took about 120 minutes and 60 minutes for air flow velocities of 0.2 m/s and 0.4 m/s, respectively. It is probably a consequence of the increased air circulation rate resulting in the air coming in contact with the activate carbon more frequently. In general, the experimental results have demonstrated the capability of the proposed airpurification process to provide and sustain good indoor air quality.
5.3 Effect of outdoor-air fraction on air handling performance
The developed mathematical model has been employed to investigate theoretically the air 14
treatment process through the cooling coil. Table 3 indicates the specifications and the operational parameters of the cooling coil. In this study, the outdoor air condition is based on Singapore’s climate which is a representative of tropical rainforest climate. The mean dry bulb temperature of outdoor air is around 31 °C, and the average relative humidity is 84%. The indoor condition is maintained at a dry bulb temperature of 24 °C and a relative humidity of 60%. Fig. 10 illustrates the influence due to changing outdoor-air fraction on the outlet air condition of the cooling coil. For a specific chilled water supply temperature, a lower outlet air temperature can be effectively achieved by reducing the outdoor-air fraction. For example, the reduction of the outlet air temperature is almost 2°C when the outdoor-air fraction is reduced from 50% to 10%. In addition, the outlet air humidity ratio decreases with lower outdoor-air fractions. It is the result of a smaller cooling load for outdoor-air. Fig. 10 also shows that the increase of 1 °C of the chilled water supply temperature leads to a rise in the outlet air temperature of approximately 0.6 °C. Similarly, the average increase of the outlet air humidity ratio was around 0.00033 kg/kg for every 1 °C increase of the chilled water supply temperature. Therefore, it can be inferred from the figure that a higher chilled water inlet temperature can be adopted by reducing the outdoor-air fraction without compromising the supply air cooling capacity. Fig. 11 further shows the required chilled water supply temperature for the cooling coil to obtain a specific outlet air temperature under varying outdoor air fractions. For example, to maintain a constant outlet air temperature of 16 °C, it is possible to increase the chilled water supply temperature from 8.3 °C to 10.8 °C when the outdoor-air fraction is regulated from 40% to 10%.
15
5.4 Chiller performance analysis
The chilled water supply temperature markedly influences the energy consumption of the vapor compression chiller. Experiments were carried out on a water-cooled scroll-type chiller with cooling capacity of 15 kW. The inlet water temperature was regulated using two water mixing chambers for both the evaporator and the condenser. The chilled water supply temperature was varied from 7 to 13 °C, while the temperature of cooling water flowing into the condenser was maintained at 30 °C. Fig. 12 presents the correlation between the chiller’s COP and the chilled water supply temperature. It is apparent that raising the chilled water supply temperature improves the chiller efficiency. As illustrated in Fig. 12, the average COP increases from 3.6 to 4.42 when the chilled water supply temperature is regulated from 7 to 13 °C. In general, the COP is improved by about 3.7% for every 1 °C increase of chilled water supply temperature. Therefore, a significant improvement on the energy efficiency of the chiller plant can be expected by employing a higher chilled water supply temperature. The energy consumption could be reduced considerably because of the improved chiller’s performance and the reduced cooling load for conditioning the outdoor air.
5.5 Energy performance evaluation
To evaluate the energy saving potential of the proposed air treatment system operating in a hot and humid climate, a typical office room in Singapore is considered. Singapore has a typical tropical climate with high temperature and humidity of small variation all year round. The office room has a total floor area of 50 m2. The maximum occupancy is 0.2 person/m2. The space cooling load is
16
around 70 W/m2. In this study, as shown in Table 4, three outdoor conditions are selected as representatives of the weather in Singapore during the daytime. When the pollutants level of outdoor air is low, the outdoor air can be used to minimize and dilute the indoor air contaminants. However, the outdoor air may at times contain unacceptably high levels of pollutants, particularly during haze situations. At such times, reducing the outdoor air flow rate and activating the air purification process should be considered to improve the indoor air quality. Fig. 13 illustrates the impact of outdoor air flow rate on the total cooling load under three outdoor conditions. As the outdoor condition is usually hot and humid in Singapore, the enthalpy of outdoor air is much higher than the indoor air. The cooling load for conditioning the outdoor airflow constitutes a large portion of the total cooling load especially for higher outdoor air flow rate. Simulation results show that the cooling load for outdoor air accounts for around 52% of the total cooling load when the outdoor air flow rate is 10 L/s per person. Once the outdoor air quality is poor, the air treatment process permits the intake of a reduced outdoor air flow rate. The outdoor air flow rate of 0 L/s per person is an extreme condition in which no outdoor air is introduced inside so that the total cooling load is only 71.8 W/m2. In general, it is observed from the figure that a lower outdoor air flow rate results in a significant reduction on the total cooling load. Fig. 14 illustrates the impact of varying outdoor air flow rate on the energy consumption for a typical outdoor condition (dry bulb temperature of outdoor air is 31 °C, and the relative humidity is 84%). The outdoor air flow rate of 10 L/s per person is required for the ventilation of a typical office [33]. In this study, the air purification process is switched on when the outdoor air flow rate is lower than the pre-set requirement. As shown in Fig. 14, a significant amount of energy can be saved from 17
operating the chiller due to the reduced outdoor air flow rate. For example, the energy consumed by chilled reduces from 41.89 W/m2 to 25.78 W/m2 when the outdoor air flow rate is regulated from 10 L/s per person to 4 L/s person. The energy saving on the chiller is a consequence of the following two factors. Firstly, the total cooling load decreases by 32% due to the reduced outdoor air flow rate. Secondly, the 11% increase in the COP of chiller is obtained by raising the chilled water temperature from 7 to 10 °C as shown in Fig. 12. Fig. 15 illustrates the total energy consumption and the energy consumption fraction of primary equipment. When compared to the conventional system of having an outdoor air flow rate of 10 L/s per person, this air treatment system enables a total energy saving of 21% by simply reducing the outdoor air flow rate to 4 L/s per person. On the flip-side, the air purification process entails a higher fan power. The pressure drop through the activated carbon filter is calculated based on Eq. (26). The motor efficiency is assumed to be 0.85 while the blower efficiency is assumed to be 0.65 [34,35] in order to compute the input power to the blower’s motor. In addition, the ozone generator consumes a power of 20W to produce ozone of 1 g/h. The reduced outdoor air intake translates to a higher flow rate for the air purification process resulting in an energy increase of the air treatment system. As illustrated in Fig. 14, when the outdoor air flow rate is regulated from 10 L/s per person to 0 L/s per person, the energy consumed by the air treatment system (fan and ozone generator) increases from 10.29 W/m2 to 16.52 W/m2. Therefore, the reduced outdoor air intake contributes to a larger energy consumption fraction for both fan and ozone generator. In Fig. 15, for example, by regulating the outdoor air flow rate from 10 L/s per person to 4 L/s per person, the energy consumption fraction of the fan increases from 19.72% to 32.81%. Fig. 15(e) further demonstrates the total energy consumption of the system under different outdoor air 18
flow rates. In sum, the overall energy saving of this air treatment system is up to 36%.
6. Conclusions
We have introduced an air treatment system which is able to employ a lower outdoor ventilation rate due to the activation of an air-purification process. A mathematical model for the cooling coil unit has been developed and validated against experimental data. Experimental studies have been carried out to investigate the performance of the air-purification and the chiller. Experimental results clearly demonstrate the capability of the proposed air-purification process to maintain an acceptable indoor air quality. The air purification function in this air treatment system allows a reduced outdoor air intake which results in a significant energy savings in hot and humid climate.
Acknowledgements
The authors gratefully acknowledge the generous funding from the National Research Foundation (NRF) Singapore under the Energy Innovation Research Programme (EIRP) Funding Scheme (R-265-000-515-279) managed on behalf by Building and Construction Authority (BCA).
19
References
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bed columns for carbon dioxide adsorption, Chemical Engineering Research and Design. 92 (2014) 961–988. doi:10.1016/j.cherd.2013.08.018. [32] X. Zhou, J. Braun, A Simplified Dynamic Model for Chilled-Water Cooling and Dehumidifying Coils—Part 2: Experimental Validation (RP-1194), HVAC&R Research. 13 (2007) 805–817. doi:10.1080/10789669.2007.10390987. [33] ASHRAE, ASHRAE Handbook-Fundamentals, Atalanta, GA, 2009. [34] ASHRAE, ASHRAE Handbook-HVAC Systems and Equipment, Atalanta, GA, 2008. [35] C.A. Roulet, F.D. Heidt, F. Foradini, M.C. Pibiri, Real heat recovery with air handling units, Energy and Buildings. 33 (2001) 495–502. doi:10.1016/S0378-7788(00)00104-3.
23
Figure captions Fig. 1. Schematic of the air treatment system. Fig. 2. Computational element for one segment of the cooling coil. Fig. 3. Experimental setup for ozone-based oxidation process. (a) Schematic diagram; (b) Photograph of the experimental setup. Fig. 4. Experimental setup for CO2 scrubbing process. (a) Schematic diagram; (b) Photograph of the experimental setup. Fig. 5. Experimental setup for the chiller. (a) Schematic diagram; (b) Photograph of the experimental setup. Fig. 6. Framework of this study. Fig. 7. Validation of the simulation results with the experimental data. (a) Outlet air temperature. (b) Outlet air humidity ratio. Fig. 8. Experimental results on the effect of ozone for VOCs from different sources. (a) VOCs from paint; (b) VOCs from glue; (c) Ethylbenzene; (d) Xylene. Fig. 9. Experimental results on CO2 reduction percentage under varying air flow velocities. Fig. 10. Outlet air condition of the cooling coil under varying outdoor-air fraction and chilled water temperature. (a) Outlet air temperature; (b) outlet air humidity ratio. Fig. 11. Effect of the outdoor-air fraction on the chilled water supply temperature. Fig. 12. Effect of the chilled water temperature on the COP of the chiller. Fig. 13. Total cooling load under varying outdoor air flow rate. 24
Fig. 14. Impact of outdoor air flow rate on the energy consumption. Fig. 15. Energy consumption fraction of the primary equipment under varying outdoor air flow rate. (a) 10 L/s per person; (b) 8 L/s per person; (c) 4 L/s per person; (d) 0 L/s per person; and (e) total energy consumption per floor area.
25
Fig. 1. Schematic of the air treatment system.
26
Fig. 2. Computational element for one segment of the cooling coil.
27
Fig. 3. Experimental setup for ozone-based oxidation process. (a) Schematic diagram; (b) Photograph of the experimental setup.
28
Fig. 4. Experimental setup for CO2 scrubbing process. (a) Schematic diagram; (b) Photograph of the experimental setup.
29
Fig. 5. Experimental setup for the chiller. (a) Schematic diagram; (b) Photograph of the experimental setup.
30
Fig. 6. Framework of this study.
31
Fig. 7. Validation of the simulation results with the experimental data. (a) Outlet air temperature. (b) Outlet air humidity ratio.
32
Fig. 8. Experimental results on the effect of ozone for VOCs from different sources. (a) VOCs from paint; (b) VOCs from glue; (c) Ethylbenzene; (d) Xylene.
33
Fig. 9. Experimental results on CO2 reduction percentage under varying air flow velocities.
34
Fig. 10. Outlet air condition of the cooling coil under varying outdoor-air fraction and chilled water temperature. (a) Outlet air temperature; (b) outlet air humidity ratio.
35
Fig. 11. Effect of the outdoor-air fraction on the chilled water supply temperature.
36
Fig. 12. Effect of the chilled water temperature on the COP of the chiller.
37
Fig. 13. Total cooling load under varying outdoor air flow rate.
38
Fig. 14. Impact of outdoor air flow rate on the energy consumption.
39
Fig. 15. Energy consumption fraction of the primary equipment under varying outdoor air flow rate. (a) 10 L/s per person; (b) 8 L/s per person; (c) 4 L/s per person; (d) 0 L/s per person; and (e) total energy consumption per floor area.
40
Table captions Table 1 The measured parameters and the specification of measuring instruments Parameter
Measurement
Accuracy
ranges VOC
0.1 to 1000 ppm
±2% of reading
0 to 10 ppm
±(0.01 ppm + 7.5% of reading)
0 to 2000 ppm
±(1.5% of range + 2% of
concentration Ozone concentration CO2 concentration
reading) Air velocity
0 to 8 m/s
±5% of reading
41
Table 2 Boundary conditions for steady-state operations Operation
Coil
Ta,in
RHa,in
ma
mw
Tw,in
(°C)
(%)
(kg/s)
(kg/s)
(°C)
Four-
28.68
55.6
1.16
0.45
2.04
Four-
28.68
55.6
0.72
0.45
2.04
Four-
25.99
70.22
1.2
0.46
1.91
Four-
25.99
54.13
1.2
0.46
1.91
Eight-
23.94
70.8
1.13
0.43
2.33
Eight-
24.59
55.02
1.27
0.24
1.87
Eight-
24.59
55.02
1.27
0.44
1.87
Eight-
27.29
51.19
1.04
0.52
3.28
condition 1 Row 2 Row 3 Row 4 Row 5 Row 6 Row 7 Row 8 Row
42
Table 3 Specifications of the cooling coil
Number of row
Number of passes
Coil length,
Longitudinal pitch, m
Transverse pitch, m
m 6
8
1.2192
Tube outer diameter, m
0.033
43
0.0381
0.0127
Table 4 Selected outdoor conditions in Singapore
Outdoor Condition
Condition 1
Dry-bulb
2
Condition 3
°C
temperature Relative humidity
Condition
%
33
31
29
75
84
90
44
S0378-7788(16)30945-8 http://dx.doi.org/doi:10.1016/j.enbuild.2017.06.059 ENB 7722
To appear in:
ENB
Received date: Revised date: Accepted date:
24-9-2016 14-2-2017 21-6-2017
Please cite this article as: X.Cui, B.Mohan, M.R.Islam, K.J.Chua, Investigating the energy performance of an air treatment incorporated cooling system for hot and humid climate, Energy and Buildingshttp://dx.doi.org/10.1016/j.enbuild.2017.06.059 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.
Investigating the energy performance of an air treatment incorporated cooling system for hot and humid climate X. Cui1, B. Mohan 1, M.R. Islam1,2, K.J. Chua1,2 * 1
Department of Mechanical Engineering, National University of Singapore, 9 Engineering Drive 1, Singapore 117576, Singapore 2 Engineering Science Programme, National University of Singapore, 9 Engineering Drive 1, Singapore 117575, Singapore *Tel. + 65 6516 2558; fax: + 65 6779 1459. E-mail address: [email protected] (K.J. Chua)
Highlights
An air treatment system has been introduced for enhanced indoor air quality.
A lower outdoor air flow rate is employed due to the air-purification process.
A mathematical model for the cooling coil unit has been developed.
Experimental systems have been constructed to study the real-time performance.
The system could achieve marked energy savings in hot and humid climate.
Abstract: An air treatment system incorporating an air-purification unit has been proposed to reduce the energy consumption of air-conditioning system and improve indoor air quality. This system employs a lower outdoor-air fraction to realize the acceptable indoor air quality and thermal comfort resulting in marked energy savings due to the reduced cooling load for the outdoor air. The performance of the primary equipment of the air treatment system has been individually investigated. Experimental systems have been constructed to study the air-purification process and to evaluate the chiller’s performance. Experimental results have demonstrated that the proposed air-purification unit is able to remove indoor air pollutants such as volatile organic compound. A computational model has concurrently been developed to determine the performance of the air handling cooling coil which is
impacted by the outdoor-air fraction and the chilled water temperature. A higher chilled water supply temperature is adopted by reducing the outdoor-air fraction without compromising the supply air cooling capacity. Having a reduced outdoor air cooling load and a better chiller’s energy efficiency operating with higher chilled water temperatures, the achievable energy savings of the airconditioning system has been demonstrated to be up to 36% under a tropical hot and humid climate.
Keywords: Air-conditioning; Heat transfer; Mathematical model; Energy consumption.
Nomenclature A cp D dp Ga h hm i H k Lef m P Pr Q Re s Sl St T u V Wf μ
ν ρ 𝜖 𝜂𝑠
area [m2] specific heat [kJ/(kg °C)] tube diameter [m] particle mean diameter air flow rate per square meter [kg/(s· m2)] convection heat transfer coefficient [kW/(m2 °C)] mass transfer coefficient [kg/(m2 s)] enthalpy of air [kJ/kg] length of the adsorption column [m] thermal conductivity [W/(m·°C)] Lewis factor mass flow rate [kg/s] pressure [kPa] Prandtl number heat transfer rate [kW] Reynolds number fin spacing [m] tube spacing in air flow direction [m] tube spacing normal to air flow [m] temperature [°C] velocity [m/s] volume flow rate of air [m3/s] fan power [kW] dynamic viscosity coefficient [Pa·s] kinematic viscosity [m2/s] density [kg/m3] void fraction surface effectiveness 2
ω
humidity ratio [kg/kg]
Subscript a air dew dew-point temperature of air i inner surface in inlet of cooling coil m mean value o outer surface out outlet of cooling coil s surface w water
1. Introduction
Outdoor-air ventilation is a common method to maintain a healthy indoor air. Increasing the ventilation rate of the air-conditioning system through greater outdoor air-intake has a positive impact on reducing sick-building syndrome (SBS) symptoms [1,2]. However, increasing the outdoor air fraction more often than not brings about greater energy consumption attributed to a higher cooling load for a higher level of fresh outdoor air, especially in hot and humid climates [3]. Therefore, it is an ongoing research challenge to find alternative cooling and ventilation strategies to reduce energy consumption and improve indoor air quality. To achieve energy efficient air conditioning systems, researchers have conducted extensive works in terms of designing and engineering novel systems [4–7]. Lin et al. [8] compared year-round energy consumptions of three ventilation methods for typical indoor spaces in Hong Kong. It was observed that the stratum ventilation provided greater energy savings when compared with mixing ventilation and displacement ventilation. Fong et al. [9] further illustrated that the rear-middle-level-
3
exhausted stratum ventilation provided a satisfied thermal comfort condition with least energy consumption. Chua et al. [10] conducted a comparison study on the dehumidification performances of three temperature control strategies, namely, chilled water flow control, bypass air control and the variable air volume control. Theoretical analysis on the cooling coil was developed to investigate the part-load performance. Niu et al. [11] proposed an air-conditioning system that combined chilledceiling with desiccant cooling for hot and humid climates. Their results highlighted that the combined system was able to save up to 44% of primary energy consumption compared with a conventional constant volume all-air system. Another low-energy air-conditioning strategy was reported to combine the employment of chilled-ceiling with the use of microencapsulated phase change material slurry storage and evaporative cooling technology to promote energy savings [12]. To curb energy consumption, a number of researchers have proposed combining the conventional airconditioning system with energy-efficient devices[13–15]. Younis et al. [16] investigated the performance of a displacement ventilation system combined with an evaporative cooled ceiling. Kim et al. [17] employed liquid desiccant dehumidification devices and evaporative coolers to enhance energy efficiency. Lee et al. [18] conducted a theoretical study on the decoupling sensible and latent cooling for desirable indoor air conditions. The results indicated that the annual energy consumption could be reduced by 54% over conventional system. In general, the energy-efficient systems reviewed relied on fresh outdoor air to maintain a desired indoor air quality. However, it may be undesirable to sustain a constant level of outdoor air intake if the outdoor air quality is poor. When such a situation occurs, the use of an air purification process is imperative to continue to sustain good indoor air quality without penalizing the energy 4
efficiency of the air-conditioning system. In modern buildings, volatile organic compound (VOC) is known as one of the primary indoor airborne pollutants [19]. It has been reported that some of VOCs may cause potential adverse health issues [20,21]. Several studies have proposed possible strategies for removing VOCs. Mo et al. [22] developed a general model for analyzing VOC removal performance of photocatalytic oxidation reactors. Cheng et al. [23] compared several building materials in terms of their VOC emissions. Building materials with mineral content obtained generated the least byproducts after the reaction with ozone. Waring and Wells [24] studied the impact of using ozone, hydroxyl radical, and the nitrate radical on indoor residential VOC conversion. They indicated that the total VOC conversion was dominated by ozonolysis and the hydroxyl radical reactions. Sidheswaran et al. [25] investigated the air cleaning performance of using activated carbon fiber (ACF) filters in air-conditioning systems. The ACF media was able to adsorb VOCs from indoor air. During the regeneration process, VOCs can be desorbed from the ACF media so that it can be used in the next cycle of air cleaning. Thus far, existing works have yet to adequately evaluate the energy performance of an airconditioning system equipped with an air-purification unit for minimizing the outdoor air fraction in hot and humid climate. In addition, in recent years, Singapore has been suffering from severe smoke haze regularly due to the numerous forest fires in Southeast Asian. The haze effect is particularly prominent during the Southwest Monsoon Season [26]. This situation further spurs us to developing innovative methods to improve energy efficiency of existing air-conditioning systems while reducing the polluted outdoor air intake and maintaining an acceptable indoor air quality. To address these issues, the present work proposes an air treatment system with an air5
purification unit for reduced cooling load and improved air quality in hot and humid climate like Singapore. The key objectives of this work are to (1) investigate the performance of individual primary equipment of an innovative air treatment system that employs ozone, and (2) evaluate the energy performance of the air-conditioning system with regulated outdoor air flow rate. We will first introduce the air-treatment system design. Thereafter, a mathematical modelling of the air conditioning system’s chilled water cooling coil will be developed, followed by the description of experimental setup that incorporates key components. Based on the performance of the air handling unit and the incorporated ozone air purifier, we further illustrate varying levels of reducing energy consumption of the air-conditioning system.
2. Description of the air treatment system
Fig. 1 presents the schematic of the proposed ozone-based air treatment system to improve air quality and to reduce energy consumption. The return air is first filtered and purified in an ozonebased oxidation device. The ozone is produced from the generator using oxygen which is generated by the combined electrolysis and photocatalysis processes. The oxygen can be used to enrich the supply air as well. Thereafter, the CO2 scrubber, filled with activated carbon filter, reduces the CO2 concentration and further purifies the supply air. Finally, the air is cooled and dehumidified in air handling unit using chilled water. Due to the air-purification and oxygen enrichment processes, this system is able to employ a lower outdoor-air fraction resulting in a significant saving on energy consumption of the air conditioning system. 6
3. Mathematical formulation
3.1 Cooling coil of air handling unit
The chilled water coil of air handling unit is an important part of the overall cooling system to maintain the comfortable indoor thermal condition. The supplied air is cooled and dehumidified by passing it through the cooling coil. To precisely determine the performance of the cooling coil, a mathematical model is developed and judiciously described in this section. The calculation is carried out using the “row-by-row method” [27,28]. The computational domain is discretized into N segments for each row of the cooling coil. Fig. 2 illustrates the computational element for one segment in which governing equations are established. For the air side, the total heat transfer can be evaluated by the enthalpy change: ∆𝑄(𝑖,𝑗) = 𝑚𝑎 (𝑖𝑎(𝑖,𝑗) − 𝑖𝑎(𝑖+1,𝑗) )
(1)
In addition, the total heat transfer is a sum of the sensible heat transfer and the latent heat transfer: ∆𝑄(𝑖,𝑗) = ∆𝑄𝑠(𝑖,𝑗) + ∆𝑄𝑙(𝑖,𝑗) = ℎ𝑜 𝛥𝐴𝑜 𝜂𝑠 (𝑇𝑎,𝑚(𝑖,𝑗) − 𝑇𝑠,𝑚(𝑖,𝑗) ) + ℎ𝑓𝑔 ℎ𝑚 𝛥𝐴𝑜 𝜂𝑠 (𝜔𝑎,𝑚(𝑖,𝑗) − 𝜔𝑠,𝑚(𝑖,𝑗) )
(2)
where 𝛥𝐴𝑜 is the outer surface area of cooling coil, and 𝜂𝑠 is the surface effectiveness. The external heat transfer coefficient can be obtained based on the Colburn j-factor analogy [29]: ℎ𝑜 = 𝑗𝐺𝑎 𝑐𝑝𝑎 𝑃𝑟 −2/3
7
(3)
where 𝑆
𝑗 = 0.14𝑅𝑒𝑎−0.328 ( 𝑡)
−0.502
𝑆𝑙
𝑠
( )
0.0312
(4)
𝐷𝑜
The Lewis factor is approximately unity for air/water mixtures 𝐿𝑒𝑓 =
ℎ𝑜
≈1
(5)
ℎ𝑜 Δ𝐴𝑜 (𝑖𝑎,𝑚(𝑖,𝑗) − 𝑖𝑠,𝑚(𝑖,𝑗) )
(6)
ℎ𝑚 ∙𝑐𝑝𝑎
Therefore, Eq. (2) can be rewritten as: ∆𝑄(𝑖,𝑗) =
𝜂𝑠 𝑐𝑝𝑎
For the chilled water, the total heat transfer can be expressed similarly as: ∆𝑄(𝑖,𝑗) = 𝑚𝑤 𝑐𝑝𝑤 (𝑇𝑤(𝑖,𝑗+1) − 𝑇𝑤(𝑖,𝑗) )
(7)
∆𝑄(𝑖,𝑗) = ℎ𝑖 ∆𝐴𝑖 (𝑇𝑠,𝑚(𝑖,𝑗) − 𝑇𝑤,𝑚(𝑖,𝑗) )
(8)
The inner heat transfer coefficient for the chilled water side can be determined by the correlation described by Petukhov [30]: 2
𝑓
𝑁𝑢𝑖 =
(8)(𝑅𝑒𝑤 −1000)𝑃𝑟𝑤
2 𝑓 0.5 3 −1) 1+12.7(8) (𝑃𝑟𝑤
[1 +
𝐷 3 ( 𝑖) ] 𝐿
(9)
where 𝑓 = (0.79 𝑙𝑛 𝑅𝑒𝑤 − 1.64)−2
(10)
𝑖𝑎,𝑚(𝑖,𝑗) , 𝑖𝑠,𝑚(𝑖,𝑗) 𝑇𝑎,𝑚(𝑖,𝑗) , 𝑇𝑠,𝑚(𝑖,𝑗) , and 𝑇𝑤,𝑚(𝑖,𝑗) are the mean value between adjacent grid point. For example, 𝑖𝑎,𝑚(𝑖,𝑗) can be expressed as 𝑖𝑎,𝑚(𝑖,𝑗) =
(𝑖𝑎(𝑖,𝑗) +𝑖𝑎(𝑖+1,𝑗) ) 2
(11)
Eliminating 𝑖𝑎(𝑖+1,𝑗) by substituting 𝑖𝑎,𝑚(𝑖,𝑗) into Eq. (1), the energy equations become: ∆𝑄(𝑖,𝑗) = 2𝑚𝑎 (𝑖𝑎(𝑖,𝑗) − 𝑖𝑎,𝑚(𝑖,𝑗) ) 8
(12)
Similarly, eliminating 𝑇𝑤(𝑖+1,𝑗) by substituting 𝑇𝑤,𝑚(𝑖,𝑗) into Eq. (7): ∆𝑄(𝑖,𝑗) = 2𝑚𝑤 𝑐𝑝𝑤 (𝑇𝑤,𝑚(𝑖,𝑗) − 𝑇𝑤(𝑖,𝑗) )
(13)
Eliminating 𝑖𝑎,𝑚(𝑖,𝑗) between Eq. (6) and Eq. (12) yields ∆𝑄(𝑖,𝑗) =
𝜂𝑠 ℎ ∆𝐴𝑜 𝑐𝑝𝑚 𝑜
∙ (𝑖𝑎(𝑖,𝑗) − 𝑖𝑠,𝑚(𝑖,𝑗) )
1+∆𝑁𝑇𝑈𝑜 /2
(14)
where ∆𝑁𝑇𝑈𝑜 =
𝜂𝑠 ℎ𝑜 ∆𝐴𝑜 𝑚𝑎 𝑐𝑝𝑚
(15)
Again, eliminating 𝑇𝑤,𝑚(𝑖,𝑗) between Eq. (8) and Eq. (13) yields ∆𝑄(𝑖,𝑗) =
ℎ𝑖 ∆𝐴𝑖 1+∆𝑁𝑇𝑈𝑖 /2
∙ (𝑇𝑠,𝑚(𝑖,𝑗) − 𝑇𝑤(𝑖,𝑗) )
(16)
where ∆𝑁𝑇𝑈𝑖 =
ℎ𝑖 ∆𝐴𝑖 𝑚𝑤 𝑐𝑝𝑤
(17)
The air enthalpy on the coil surface (𝑖𝑠,𝑚(𝑖,𝑗) ) and the coil surface temperature (𝑇𝑠,𝑚(𝑖,𝑗) ) are essential parameters for calculating the ultimate temperature and humidity ratio of air. The coil surface condition depends largely on its surface temperature. When the coil surface temperature is higher than the dew point temperature of the air, the surface of the cooling coil remains dry so that the relationship between 𝑖𝑠,𝑚(𝑖,𝑗) and 𝑇𝑠,𝑚(𝑖,𝑗) is given as 𝑖𝑠,𝑚(𝑖,𝑗) = 𝑖𝑎(𝑖,𝑗) + 𝑐𝑝𝑎 (𝑇𝑠,𝑚(𝑖,𝑗) − 𝑇𝑎(𝑖,𝑗) ), 𝑇𝑠,𝑚(𝑖,𝑗) > 𝑇𝑑𝑒𝑤(𝑖,𝑗)
(18)
When the coil surface temperature is lower than the dew point temperature of the air, the coil surface is wet since the air is dehumidified by passing through the coil. In such case, 𝑖𝑠,𝑚(𝑖,𝑗) should
9
be evaluated by the enthalpy of the saturated air at the coil outer surface temperature as follows: 2 𝑖𝑠,𝑚(𝑖,𝑗) = 10.76 + 1.4𝑇𝑠,𝑚(𝑖,𝑗) + 0.046𝑇𝑠,𝑚(𝑖,𝑗) , 𝑇𝑠,𝑚(𝑖,𝑗) ≤ 𝑇𝑑𝑒𝑤(𝑖,𝑗)
(19)
After the determination of 𝑇𝑠,𝑚(𝑖,𝑗) , the total heat transfer ∆𝑄(𝑖,𝑗) can be calculated using Eq. (16). Thereafter, the parameters for the following grid can be calculated as 𝑖𝑎(𝑖+1,𝑗) = 𝑖𝑎(𝑖,𝑗) − 𝑇𝑤(𝑖,𝑗+1) = 𝑇𝑤(𝑖,𝑗) −
∆𝑄(𝑖,𝑗)
(20)
𝑚𝑎 ∆𝑄(𝑖,𝑗)
(21)
𝑚𝑤 𝑐𝑝𝑤
In addition, by solving sensible heat transfer equations, we obtain 𝑇𝑎(𝑖+1,𝑗) =
∆𝑁𝑇𝑈𝑜 2 ∆𝑁𝑇𝑈 1+ 2 𝑜
1−
𝜔𝑎(𝑖+1,𝑗) =
∙ 𝑇𝑎(𝑖,𝑗) +
∆𝑁𝑇𝑈𝑜
∆𝑁𝑇𝑈𝑜 2
1+
∙ 𝑇𝑠,𝑚(𝑖,𝑗)
𝑖𝑎(𝑖+1,𝑗) −𝑐𝑝𝑎 ∙𝑇𝑎(𝑖+1,𝑗) [2501+1.89∙𝑇𝑎(𝑖+1,𝑗) ]∙1000
(22) (23)
3.2 Energy consumption of primary equipment
The electrical power consumed by the chiller is provided below as 𝑐ℎ𝑖𝑙𝑙𝑒𝑟 𝑝𝑜𝑤𝑒𝑟 =
𝑄𝑐 𝐶𝑂𝑃
(24)
where 𝑄𝑐 is the cooling load, and COP is the coefficient of performance of the chiller. The fan performance is calculated based on its operational air flow rate, pressure rise and efficiency. The total power is given as: 𝑊𝑓 =
𝑉∆𝑝 𝜂𝑏 𝜂𝑚
where 𝜂𝑏 is the blower efficiency, and 𝜂𝑚 is the motor efficiency. 10
(25)
The pressure drop through the activated carbon particles can be calculated as [31]: ∆𝑝 =
150𝜇𝐻 (1−𝜖)2 2 𝑑𝑝
𝜖3
𝑢+
1.75𝐿𝜌 (1−𝜖) 𝑑𝑝
𝜖3
𝑢2
(26)
where 𝑑𝑝 is the particle mean diameter, 𝜖 is the void fraction, and u is the air velocity.
4. Experimental setup
4.1 Air purification process
To independently evaluate the air purification performance of the ozone-based oxidation process, a separate experimental setup was fabricated as shown in Fig. 3. The experimental setup included an insulated air-tight chamber, an ozone generator, and sensors for measuring air parameters such as VOC concentration. The insulated chamber was made from stainless steel with size of 1m × 1m × 1m. A fan with variable speed drive was installed to mix the air inside the chamber. The ozone can be injected into the chamber through the ozone generator which was regulated by a mass flow controller. Sensors were inserted into the chamber to measure the concentration of various gases, such as VOC and Ozone. Table 1 indicates the specifications of measuring instruments. The following procedures are employed for testing the reaction between VOCs and ozone. Firstly, samples of building material or chemicals were placed inside the chamber for VOC emission. Secondly, the samples were then removed and the concentration of VOC was monitored until its concentration becomes relatively stable. Thirdly, a specific quantity of ozone was injected into the chamber. The ozone generator was turned off once a required ozone concentration was reached.
11
Consolidated data collected from sensors were monitored and recorded to observe the impact of ozone. Another key component of the air treatment system is the CO2 scrubbing unit. Fig. 4 depicts the experimental setup for CO2 scrubbing process. A chamber with a size of 750 mm × 750 mm × 750 mm was designed to simulate the indoor environment. The chamber was equipped with a small fan for mixing the air. The CO2 cylinder provided the CO2 gas into the chamber through a mass flow controller. The blower with variable speed drive was able to circulate the air at specific flow rate through the CO2 scrubber canister. In this study, the canister was filled with activated carbon which was used as the CO2 absorbing material. CO2 sensors monitored and recorded CO2 concentrations at various locations in order to investigate the effect of air flow velocity on CO2 scrubbing performance. The specifications of measuring instruments are shown in Table 1.
4.2 Chiller
Experiments were also carried out to analyze the performance of a chiller at different chilled water supply temperatures. Fig. 5 illustrates the schematic diagram of the experimental setup. It consisted of a water cooled scroll chiller of cooling capacity 15 kW, shell and tube type evaporator and condenser, two water mixing chambers, pumps, and a cooling tower. The inlet water temperature was regulated using two water mixing chambers for both the evaporator and the condenser. The chilled water supply temperature was varied from 7 to 13 °C. The cooling tower was installed to provide cooled water to be conditioned and circulated back to the mixing chambers.
12
5. Results and discussion
In this section, we first validate the cooling coil mathematical model, followed by the experimental results on the air purification process and the performance of chiller. Based on the simulation results and experimental data, we further evaluate the thermal performance of the air handling unit and the energy consumption of the air treatment system. The framework of this study is illustrated in Fig. 6.
5.1 Model validation on cooling coil
The mathematical modeling was validated against experimental data acquired from a previous study [32]. Zhou and Braun have investigated the performance of both four-row and eight-row cooling coils [32]. Experimental study was conducted under a variety of operating conditions for inlet air flow velocity, inlet air temperature, inlet air humidity, water flow rate, and inlet water temperature. Table 2 indicates the boundary conditions for steady-state operations. The computational model was validated by comparing the simulated outlet conditions with experimental data. Fig. 7 demonstrates the capability of the computational model to predict the cooling coil performance. It portrayed a good agreement with experiments a discrepancy of less than ±10%.
5.2 Air purification performance
VOCs are usually emitted from building materials such as paint and glue. In this study, samples 13
of paint and glue were tested for VOC abatement using ozone in the experimental chamber. In addition, two pure chemicals, namely, Ethylbenzene and Xylene, were also treated in the chamber to investigate the effect of oxidation reaction. Fig. 8 presents the experimental results on VOCs treatment. The VOC for paint and glue are reduced by around 19% and 44%, respectively, over a time period of 24 hours. Comparing with paint sample, the glue sample demonstrates a faster reaction rate and consumes more ozone within the same time period. It is also observed that the reduction rate for Ethylbenzene and Xylene are 23% and 26%, respectively, over a time period of 30 hours. The experimental result shows the feasibility of using ozone to markedly reduce VOCs in an enclosed environment. Fig. 9 further shows the CO2 concentration reduction percentage in the experimental chamber. The initial CO2 concentration in the chamber was kept as around 1200 ppm. To evaluate the performance of canister in CO2 reduction, the air was circulated through the canister under varying velocities. It can be seen that a higher air flow velocity leads to improve CO2 reduction rate. For example, to obtain 15% reduction of CO2, it took about 120 minutes and 60 minutes for air flow velocities of 0.2 m/s and 0.4 m/s, respectively. It is probably a consequence of the increased air circulation rate resulting in the air coming in contact with the activate carbon more frequently. In general, the experimental results have demonstrated the capability of the proposed airpurification process to provide and sustain good indoor air quality.
5.3 Effect of outdoor-air fraction on air handling performance
The developed mathematical model has been employed to investigate theoretically the air 14
treatment process through the cooling coil. Table 3 indicates the specifications and the operational parameters of the cooling coil. In this study, the outdoor air condition is based on Singapore’s climate which is a representative of tropical rainforest climate. The mean dry bulb temperature of outdoor air is around 31 °C, and the average relative humidity is 84%. The indoor condition is maintained at a dry bulb temperature of 24 °C and a relative humidity of 60%. Fig. 10 illustrates the influence due to changing outdoor-air fraction on the outlet air condition of the cooling coil. For a specific chilled water supply temperature, a lower outlet air temperature can be effectively achieved by reducing the outdoor-air fraction. For example, the reduction of the outlet air temperature is almost 2°C when the outdoor-air fraction is reduced from 50% to 10%. In addition, the outlet air humidity ratio decreases with lower outdoor-air fractions. It is the result of a smaller cooling load for outdoor-air. Fig. 10 also shows that the increase of 1 °C of the chilled water supply temperature leads to a rise in the outlet air temperature of approximately 0.6 °C. Similarly, the average increase of the outlet air humidity ratio was around 0.00033 kg/kg for every 1 °C increase of the chilled water supply temperature. Therefore, it can be inferred from the figure that a higher chilled water inlet temperature can be adopted by reducing the outdoor-air fraction without compromising the supply air cooling capacity. Fig. 11 further shows the required chilled water supply temperature for the cooling coil to obtain a specific outlet air temperature under varying outdoor air fractions. For example, to maintain a constant outlet air temperature of 16 °C, it is possible to increase the chilled water supply temperature from 8.3 °C to 10.8 °C when the outdoor-air fraction is regulated from 40% to 10%.
15
5.4 Chiller performance analysis
The chilled water supply temperature markedly influences the energy consumption of the vapor compression chiller. Experiments were carried out on a water-cooled scroll-type chiller with cooling capacity of 15 kW. The inlet water temperature was regulated using two water mixing chambers for both the evaporator and the condenser. The chilled water supply temperature was varied from 7 to 13 °C, while the temperature of cooling water flowing into the condenser was maintained at 30 °C. Fig. 12 presents the correlation between the chiller’s COP and the chilled water supply temperature. It is apparent that raising the chilled water supply temperature improves the chiller efficiency. As illustrated in Fig. 12, the average COP increases from 3.6 to 4.42 when the chilled water supply temperature is regulated from 7 to 13 °C. In general, the COP is improved by about 3.7% for every 1 °C increase of chilled water supply temperature. Therefore, a significant improvement on the energy efficiency of the chiller plant can be expected by employing a higher chilled water supply temperature. The energy consumption could be reduced considerably because of the improved chiller’s performance and the reduced cooling load for conditioning the outdoor air.
5.5 Energy performance evaluation
To evaluate the energy saving potential of the proposed air treatment system operating in a hot and humid climate, a typical office room in Singapore is considered. Singapore has a typical tropical climate with high temperature and humidity of small variation all year round. The office room has a total floor area of 50 m2. The maximum occupancy is 0.2 person/m2. The space cooling load is
16
around 70 W/m2. In this study, as shown in Table 4, three outdoor conditions are selected as representatives of the weather in Singapore during the daytime. When the pollutants level of outdoor air is low, the outdoor air can be used to minimize and dilute the indoor air contaminants. However, the outdoor air may at times contain unacceptably high levels of pollutants, particularly during haze situations. At such times, reducing the outdoor air flow rate and activating the air purification process should be considered to improve the indoor air quality. Fig. 13 illustrates the impact of outdoor air flow rate on the total cooling load under three outdoor conditions. As the outdoor condition is usually hot and humid in Singapore, the enthalpy of outdoor air is much higher than the indoor air. The cooling load for conditioning the outdoor airflow constitutes a large portion of the total cooling load especially for higher outdoor air flow rate. Simulation results show that the cooling load for outdoor air accounts for around 52% of the total cooling load when the outdoor air flow rate is 10 L/s per person. Once the outdoor air quality is poor, the air treatment process permits the intake of a reduced outdoor air flow rate. The outdoor air flow rate of 0 L/s per person is an extreme condition in which no outdoor air is introduced inside so that the total cooling load is only 71.8 W/m2. In general, it is observed from the figure that a lower outdoor air flow rate results in a significant reduction on the total cooling load. Fig. 14 illustrates the impact of varying outdoor air flow rate on the energy consumption for a typical outdoor condition (dry bulb temperature of outdoor air is 31 °C, and the relative humidity is 84%). The outdoor air flow rate of 10 L/s per person is required for the ventilation of a typical office [33]. In this study, the air purification process is switched on when the outdoor air flow rate is lower than the pre-set requirement. As shown in Fig. 14, a significant amount of energy can be saved from 17
operating the chiller due to the reduced outdoor air flow rate. For example, the energy consumed by chilled reduces from 41.89 W/m2 to 25.78 W/m2 when the outdoor air flow rate is regulated from 10 L/s per person to 4 L/s person. The energy saving on the chiller is a consequence of the following two factors. Firstly, the total cooling load decreases by 32% due to the reduced outdoor air flow rate. Secondly, the 11% increase in the COP of chiller is obtained by raising the chilled water temperature from 7 to 10 °C as shown in Fig. 12. Fig. 15 illustrates the total energy consumption and the energy consumption fraction of primary equipment. When compared to the conventional system of having an outdoor air flow rate of 10 L/s per person, this air treatment system enables a total energy saving of 21% by simply reducing the outdoor air flow rate to 4 L/s per person. On the flip-side, the air purification process entails a higher fan power. The pressure drop through the activated carbon filter is calculated based on Eq. (26). The motor efficiency is assumed to be 0.85 while the blower efficiency is assumed to be 0.65 [34,35] in order to compute the input power to the blower’s motor. In addition, the ozone generator consumes a power of 20W to produce ozone of 1 g/h. The reduced outdoor air intake translates to a higher flow rate for the air purification process resulting in an energy increase of the air treatment system. As illustrated in Fig. 14, when the outdoor air flow rate is regulated from 10 L/s per person to 0 L/s per person, the energy consumed by the air treatment system (fan and ozone generator) increases from 10.29 W/m2 to 16.52 W/m2. Therefore, the reduced outdoor air intake contributes to a larger energy consumption fraction for both fan and ozone generator. In Fig. 15, for example, by regulating the outdoor air flow rate from 10 L/s per person to 4 L/s per person, the energy consumption fraction of the fan increases from 19.72% to 32.81%. Fig. 15(e) further demonstrates the total energy consumption of the system under different outdoor air 18
flow rates. In sum, the overall energy saving of this air treatment system is up to 36%.
6. Conclusions
We have introduced an air treatment system which is able to employ a lower outdoor ventilation rate due to the activation of an air-purification process. A mathematical model for the cooling coil unit has been developed and validated against experimental data. Experimental studies have been carried out to investigate the performance of the air-purification and the chiller. Experimental results clearly demonstrate the capability of the proposed air-purification process to maintain an acceptable indoor air quality. The air purification function in this air treatment system allows a reduced outdoor air intake which results in a significant energy savings in hot and humid climate.
Acknowledgements
The authors gratefully acknowledge the generous funding from the National Research Foundation (NRF) Singapore under the Energy Innovation Research Programme (EIRP) Funding Scheme (R-265-000-515-279) managed on behalf by Building and Construction Authority (BCA).
19
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Figure captions Fig. 1. Schematic of the air treatment system. Fig. 2. Computational element for one segment of the cooling coil. Fig. 3. Experimental setup for ozone-based oxidation process. (a) Schematic diagram; (b) Photograph of the experimental setup. Fig. 4. Experimental setup for CO2 scrubbing process. (a) Schematic diagram; (b) Photograph of the experimental setup. Fig. 5. Experimental setup for the chiller. (a) Schematic diagram; (b) Photograph of the experimental setup. Fig. 6. Framework of this study. Fig. 7. Validation of the simulation results with the experimental data. (a) Outlet air temperature. (b) Outlet air humidity ratio. Fig. 8. Experimental results on the effect of ozone for VOCs from different sources. (a) VOCs from paint; (b) VOCs from glue; (c) Ethylbenzene; (d) Xylene. Fig. 9. Experimental results on CO2 reduction percentage under varying air flow velocities. Fig. 10. Outlet air condition of the cooling coil under varying outdoor-air fraction and chilled water temperature. (a) Outlet air temperature; (b) outlet air humidity ratio. Fig. 11. Effect of the outdoor-air fraction on the chilled water supply temperature. Fig. 12. Effect of the chilled water temperature on the COP of the chiller. Fig. 13. Total cooling load under varying outdoor air flow rate. 24
Fig. 14. Impact of outdoor air flow rate on the energy consumption. Fig. 15. Energy consumption fraction of the primary equipment under varying outdoor air flow rate. (a) 10 L/s per person; (b) 8 L/s per person; (c) 4 L/s per person; (d) 0 L/s per person; and (e) total energy consumption per floor area.
25
Fig. 1. Schematic of the air treatment system.
26
Fig. 2. Computational element for one segment of the cooling coil.
27
Fig. 3. Experimental setup for ozone-based oxidation process. (a) Schematic diagram; (b) Photograph of the experimental setup.
28
Fig. 4. Experimental setup for CO2 scrubbing process. (a) Schematic diagram; (b) Photograph of the experimental setup.
29
Fig. 5. Experimental setup for the chiller. (a) Schematic diagram; (b) Photograph of the experimental setup.
30
Fig. 6. Framework of this study.
31
Fig. 7. Validation of the simulation results with the experimental data. (a) Outlet air temperature. (b) Outlet air humidity ratio.
32
Fig. 8. Experimental results on the effect of ozone for VOCs from different sources. (a) VOCs from paint; (b) VOCs from glue; (c) Ethylbenzene; (d) Xylene.
33
Fig. 9. Experimental results on CO2 reduction percentage under varying air flow velocities.
34
Fig. 10. Outlet air condition of the cooling coil under varying outdoor-air fraction and chilled water temperature. (a) Outlet air temperature; (b) outlet air humidity ratio.
35
Fig. 11. Effect of the outdoor-air fraction on the chilled water supply temperature.
36
Fig. 12. Effect of the chilled water temperature on the COP of the chiller.
37
Fig. 13. Total cooling load under varying outdoor air flow rate.
38
Fig. 14. Impact of outdoor air flow rate on the energy consumption.
39
Fig. 15. Energy consumption fraction of the primary equipment under varying outdoor air flow rate. (a) 10 L/s per person; (b) 8 L/s per person; (c) 4 L/s per person; (d) 0 L/s per person; and (e) total energy consumption per floor area.
40
Table captions Table 1 The measured parameters and the specification of measuring instruments Parameter
Measurement
Accuracy
ranges VOC
0.1 to 1000 ppm
±2% of reading
0 to 10 ppm
±(0.01 ppm + 7.5% of reading)
0 to 2000 ppm
±(1.5% of range + 2% of
concentration Ozone concentration CO2 concentration
reading) Air velocity
0 to 8 m/s
±5% of reading
41
Table 2 Boundary conditions for steady-state operations Operation
Coil
Ta,in
RHa,in
ma
mw
Tw,in
(°C)
(%)
(kg/s)
(kg/s)
(°C)
Four-
28.68
55.6
1.16
0.45
2.04
Four-
28.68
55.6
0.72
0.45
2.04
Four-
25.99
70.22
1.2
0.46
1.91
Four-
25.99
54.13
1.2
0.46
1.91
Eight-
23.94
70.8
1.13
0.43
2.33
Eight-
24.59
55.02
1.27
0.24
1.87
Eight-
24.59
55.02
1.27
0.44
1.87
Eight-
27.29
51.19
1.04
0.52
3.28
condition 1 Row 2 Row 3 Row 4 Row 5 Row 6 Row 7 Row 8 Row
42
Table 3 Specifications of the cooling coil
Number of row
Number of passes
Coil length,
Longitudinal pitch, m
Transverse pitch, m
m 6
8
1.2192
Tube outer diameter, m
0.033
43
0.0381
0.0127
Table 4 Selected outdoor conditions in Singapore
Outdoor Condition
Condition 1
Dry-bulb
2
Condition 3
°C
temperature Relative humidity
Condition
%
33
31
29
75
84
90
44