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Energy and Buildings 43 (2011) 130–136

Contents lists available at ScienceDirect

Energy and Buildings

journa l homepage: www.e lsev ier .com/ locate /enbui ld

omparison of annual energy performances with different ventilation methodsor cooling

hang Lin ∗, C.K. Lee, Square Fong, T.T. Chow, Ting Yao, A.L.S. Chanuilding Energy & Environmental Technology Research Unit, School of Energy and Environment & Division of Building Science and Technology,ity University of Hong Kong, 83 Tat Chee Avenue, Hong Kong SAR, China

r t i c l e i n f o

rticle history:eceived 3 August 2010eceived in revised form 21 August 2010ccepted 28 August 2010

a b s t r a c t

Stratum ventilation has been proposed to cope for elevated indoor temperatures recommended by gov-ernments in East Asia. TRNSYS is used for computation of the space cooling load and system energyconsumption. Typical configurations of an office, a classroom and a retail shop in Hong Kong are inves-tigated. Compared with mixing ventilation and displacement ventilation, stratum ventilation derives its

eywords:tratum ventilationisplacement ventilationixing ventilation

RNSYS simulation

energy saving potential largely from the following three factors: the reduction in ventilation and trans-mission loads and increased COP of chillers. The year-round energy saving is found to be substantial at25% and 44% at least when compared with displacement ventilation and mixing ventilation, respectively.

© 2010 Elsevier B.V. All rights reserved.

ear-roundnergy saving

. Introduction

Minimizing the energy consumption by air-conditioning sys-ems would help to reduce carbon emission. Proactive actionsn this regard have been taken by several governments in Eastsia [1–5]. A survey shows that many facility professionals aredjusting the thermostat to higher settings in the summer to cutnergy consumption in United States [6]. In order to cope with theecommended elevated room temperatures recommended by theovernmental guidelines, stratum ventilation, a new ventilationethod, was proposed by Lin et al. with the following characteris-

ics [7,8]:

. reverse temperature gradient in the occupied zone;

. higher air speed at the head–chest level for equal air supplyvolume;

. higher supply air temperature; and

. higher evaporating temperature for the associated refrigerationplant, thus higher coefficient of performance (COP) for the refrig-eration machine(s).

Lin et al. developed an RNG k–ε model based on the commercialrogram of FUENT to perform CFD analysis. This model has beenalidated for stratum ventilation [9]. Tian et al. investigated the

∗ Corresponding author.E-mail address: bsjzl@cityu.edu.hk (Z. Lin).

378-7788/$ – see front matter © 2010 Elsevier B.V. All rights reserved.oi:10.1016/j.enbuild.2010.08.033

indoor air quality and thermal comfort of an office room with stra-tum ventilation using a numerical method [10]. It was shown that, ifreasonably designed, stratum ventilation was able to provide goodindoor air quality in the breathing zone and to achieve good thermalcomfort measured by PMV and PPD. Tian et al. investigated the dif-fusion of CO2 released by the manikin and thermal comfort understratum ventilation experimentally [11]. The results demonstratedthe flow pattern formed by stratum ventilation was able to providegood IAQ in the breathing zone. Tian et al. discussed the particle dif-fusion in an office room under stratum ventilation by adopting thenumerical method validated by their experimental data and com-pared the results with those of displacement ventilation with thesame supply air parameters, heat sources and furniture arrange-ments [12]. The particle concentrations of the entire room and thebreathing zone under stratum ventilation are less than those underdisplacement ventilation, which imply that the risk of inhaling par-ticles, particularly in the breathing zone, is smaller under stratumventilation than under displacement ventilation. Although a fewaspects of stratum ventilation have been investigated as mentionedabove, our knowledge of stratum ventilation is still very limited.

Wyon [13] pointed out that about 40% of all energy is usedfor building-related activities, these include the construction ofbuildings, heating, cooling or lighting of buildings as required and

ventilation for buildings. Energy can be saved from the afore-mentioned processes if each of them is performed efficiently. Viasimulation, Ke et al. found that supply air temperature and pri-mary airflow rate were the two proper optimizable parameterson the air side of heating, ventilating, and air-conditioning (HVAC)

Z. Lin et al. / Energy and Buildi

Nomenclature

c specific heat capacity of air (kJ kg−1 K−1)COP coefficient of performanceHroom zone height (m)m supply air mass flow rate (kg s−1)Qsen sensible load of building zone (kW)T temperature (◦C)Tgrad temperature gradient along the zone height

(◦C m−1)�r dimensionless temperature coefficient as defined in

Eq. (4)

Subscriptse exhaust airfl zone air at 0.1 m above floor

siotetetcoCtbsffithds

2

tsawtpbttttdtbtw[

atw

r reference zone points supply air

ystems [14]. Zhivov and Rymkevich calculated heating and cool-ng consumption by the HVAC system throughout the year-roundperation and compared operation costs [15]. Florides et al. usedhe TRNSYS program, for the modelling and simulation of thenergy flows of modern houses, to examine measures to reducehe thermal load [16]. Controlled ventilation is one of the measuresxamined. Through hour-by-hour simulations, Niu et al. calculatedhe annual energy consumption of an HVAC system combininghilled ceiling with desiccant cooling and compared it with thatf other systems [17]. A numerical procedure, based on couplingFD simulation and dynamic cooling load simulation, is proposedo predict annual energy consumption of the under-floor air distri-ution system by Xu and Niu [18]. This system derives its energyaving potential from the following three factors: an extendedree cooling period, a reduced ventilation load, and increased coef-cients of performance (COP) for chillers. Karlsson et al. usedhree different simulation tools to simulate a low energy terracedouse. They discussed the relative importance to the annual energyemand prediction of different energy aspects, including choice ofoftware and ventilation rates [19].

. System description

Mixing ventilation, displacement ventilation and stratum ven-ilation may be used to cool a building zone. Fig. 1 shows thechematic diagram for the three systems. For mixing ventilation,ir is supplied to the building zone at high velocity at ceiling levelhich mixes fully with the room air before being extracted from

he building zone (usually at ceiling level). The air temperature andollutant concentration is assumed to be homogeneous within theuilding zone. For displacement ventilation, air is supplied at/nearhe floor level at low velocity which fills the entire floor area andhen displaces upward for extraction at ceiling level. The tempera-ure (and usually also the pollutant concentration) increases withhe height. For stratum ventilation, air is supplied at an interme-iate level close to the breathing zone/head level which keeps theemperature and pollutant concentration to a low level within thereathing zone. Consequently, a positive temperature and pollu-ant concentration gradient is established above the breathing zonehile a negative gradient is developed below the breathing zone

9].An air-handling unit, consisting of a cooling coil and a supply

ir fan, and a water-cooled vapor compression chiller are usedo provide the cooled supply air to the building zone. The chilledater entering the cooling coil is controlled by a three-way supply

ngs 43 (2011) 130–136 131

chilled water valve based on the reference zone air temperaturein the cases of displacement ventilation and stratum ventilation,or the return air temperature in the case of mixing ventilation.The air-handling unit controller generates a linear signal from 0to 1 corresponding to nil to full flow to the coil when the zonetemperature varies between the temperatures within ±1 ◦C of theset point. To avoid instability when simulating the cooling coil,the valve will open at least 30% when the output from the air-handling unit controller is greater than zero. The operation of thechiller, the condenser water pump and the cooling tower is gov-erned by a return chilled water thermostat with a dead band of3 ◦C. The cooling tower is additionally controlled by a return con-denser water thermostat triggered between 15 and 20 ◦C. To offerbetter performance for the chiller during the low-load period, apart-load controller is used to modulate the compressor speed from30 to 50 Hz in five steps through a frequency inverter (FI) in orderto maintain the return chilled water temperature at the designvalue. The frequency starts at 30 Hz when the return chilled watertemperature exceeds the design value, and steps up if the watertemperature drops less than 1 ◦C after one simulation time step.Conversely, when the return chilled water temperature is below thedesign value, the frequency steps down if the water temperaturedrops more than 1 ◦C after one simulation time step. The fresh air-flow rate is based on 0.01 m3/s per occupant. The air-handling unitand the chilled water pump will always be running within the dailyoccupancy schedule. Three building types, an office, a classroomand a retail shop, are investigated. The zone loadings are calculatedbased on the varying ambient conditions and the pre-defined inter-nal heat sources. Hence, the annual performance of the system canbe studied.

3. Model formulations

Lee presented a model for a variable-speed ground-source heatpump which is also applicable to a variable-speed chiller [20].A parametric model as used by Jin and Spitler was adopted forthe compressor where the various parameter values includingthe swept volume, volumetric coefficient, polytropic compressioncoefficient and the mechanical efficiency were determined using aparameter estimation technique based on performance data fromthe manufacturer (www.emersonclimate.com) [21]. The perfor-mance of the condenser and evaporator coils were calculated bydividing each coil into a maximum of three portions correspondingto the different states of the refrigerant in the coils. By employ-ing different temperature effectiveness formulations for the coilportions, the performance of the entire coil was then solved byan iterative method. A thermostatic expansion valve was usedwhich maintained a constant degree of superheat at the compres-sor suction. The solution for the complete chiller cycle was thendetermined based on the energy and refrigerant mass balances forthe entire cycle as proposed by Domanski and Didion [22].

For the displacement ventilation, Skistad proposed a rule-of-thumb which stated that the room temperature sufficiently distantfrom the supply air terminal at 0.1 m above the floor is approx-imately the mean between the supply and return/exhaust airtemperatures. Moreover, the room temperature increased linearlyfrom the floor to the ceiling. The reference zone temperature wasdefined at 1.1 m above the floor [23]. Hence,

Tfl = Ts + Te

2, Tr = Tfl + Tgrand (1)

where

Tgrad = Te − Tfl

Hroom − 0.1(2)

132 Z. Lin et al. / Energy and Buildings 43 (2011) 130–136

F d (c) stratum ventilation systems equipped with an air-handling unit and a variable-speedw

a

Q

b

ts

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aztuotmsda

course. The schedules for the office rooms, classroom and retailshop are determined based on the same principle. Combined withthe design loads stated in Table 1, the peak cooling load can becalculated.

Table 1Summarized design loads for various configurations.

Type Office Classroom Shop

ig. 1. Schematic diagrams of: (a) mixing ventilation, (b) displacement ventilation anater-cooled vapor compression chiller.

nd

sen = mc(Te − Ts) (3)

Eqs. (1)–(3) can be solved if the supply air-conditions and theuilding load are known.

For stratum ventilation, the correlation of the reference zoneemperature to the supply and exhaust air temperatures are repre-ented by a dimensionless temperature defined as:

r = Tr − Ts

Te − Ts(4)

Based on the experimental results from Tian, a value of 0.726as derived under a supply and exhaust air temperature difference

f 6.8 ◦C [24]. Because the conditions are not expected to deviateignificantly from the experimental conditions, the product of theimensionless temperature multiplied by the supply and exhaustir temperature difference is expected to remain fairly constant (i.e..9368) for a non-industrial air-conditioned room under slightlyeviated temperature differences. Eq. (3) also applies for stratumentilation. The same form of mathematical expression was alsosed to present experimental results both for temperature gradi-nt and for humidity ratio gradient [25]. In estimating the referenceone humidity for both displacement ventilation and stratum ven-ilation, the humidity ratio is deemed to vary in the same fashions the temperature along the room height.

. System specification

Figs. 2–4 show the configurations of the typical office, classroomnd retail shop used in this study. The office is divided into twoones, namely the individual room and the open concourse withwo air-handling units used. The west external wall of the individ-al room zone, with a window area ratio of 0.5, is 8.4 m long. Allther partitions are assumed to be thermally symmetrical, meaning

hat there is no heat transfer across those boundaries. Single-zone

odels are adopted for the classroom and the retail shop with oneide of the external wall of length 8.4 m facing the west with a win-ow area ratio of 0.5. Again, all the other partitions are treated asdiabatic.

Fig. 2. Configuration of typical office.

Design loads for the three building types are summarized inTable 1. The equipment load is assumed to be 30% radiative and70% convective. The computer load for the retail shop is includedin the equipment load.

Fig. 5 shows the daily operating schedule for the office con-

Room Concourse

Occupant 2 8 41 27Lighting (W m−2) 19 19 17.4 70Computer (W per occupant) 150 150 0 0Equipment (W) 0 1850 220 1270

Z. Lin et al. / Energy and Buildings 43 (2011) 130–136 133

Fig. 3. Configuration of typical classroom.

Fig. 4. Configuration of typical retail shop.

Table 2Summarized design requirements for the air-conditioning systems.

Case Total zone load (kW) System lo

Office rooms, mixing 3.265 4.167Office rooms, displacement 3.201 3.834Office rooms, stratum 3.137 3.510Office concourse, mixing 4.342 7.943Office concourse, displacement 4.324 6.854Office concourse, stratum 4.306 5.732Classroom, mixing 7.889 26.24Classroom, displacement 7.782 20.02Classroom, stratum 7.676 13.12Retail shop, mixing 12.03 24.15Retail shop, displacement 11.93 20.01Retail shop, stratum 11.82 15.37

Fig. 5. Daily operating schedules for office concourse.

The thermal neutral temperatures under the three ventila-tion methods are different. The design temperatures for roomsventilated with these methods should therefore not be identical.As a starting point, the design methods of displacement venti-lation and stratum ventilation consider sensible loads only—nodesign relative humidity is defined. The corresponding settings are24 ◦C/54% for mixing ventilation, 25 ◦C for displacement ventilationand 26.5 ◦C for stratum ventilation. The design outdoor conditionis 33.5 ◦C/68%. Table 2 summarizes the design cooling loads andrequired supply air-conditions for the various cases. To allow bet-ter comparison of energy consumption, the design air flow rate ofthe mixing ventilation case is also applied to the displacement ven-tilation and stratum ventilation cases. The system load includes theventilation load resulting from the fresh air intake. The zone loadsof stratum ventilation are the lowest followed by those of the dis-placement ventilation, though the differences are less than 4% whencompared with those of the mixing ventilation. This is because thehigher design indoor temperatures adopted, the less heat trans-mits through the external wall. The systems loads drop much morethan the zone loads with the adoption of displacement ventilationand stratum ventilation. This is because the exhaust air or returnair temperatures are substantially higher for displacement ventila-tion and for stratum ventilation which results in a large reductionin the ventilation loads. The effects are especially significant forthe classroom and the retail shop in which the fresh air ratios aremuch higher than that for the office. In fact, the system load for theclassroom decreases by 24% with the adoption of displacement ven-tilation and even up to 50% with the use of stratum ventilation whencompared with that for mixing ventilation. With a smaller chillercapacity, the corresponding requirements for the water pumps, thecooling tower and the cooling coil can also be reduced which means

that both carbon foot print and cost in the construction stage canbe lowered. This is also true in the operation stage.

With the system requirements fixed, the various parameters forthe chiller are to be selected. As the supply air temperatures for the

ad (kW) Supply airflow rate (m3/s) Supply air temp. (◦C)

0.259 13.970.259 18.170.259 21.690.315 13.620.315 17.810.315 21.530.424 12.110.424 17.080.424 21.620.641 12.050.641 16.990.641 21.57

134 Z. Lin et al. / Energy and Buildings 43 (2011) 130–136

Table 3Parameter values for the chiller system used in office.

Mixing Displacement Stratum

ChillerCondenser water mass flowrate (kg s−1) 0.77 0.68 0.59Chilled water mass flowrate (kg s−1) 0.58 0.51 0.44Swept volume of compressor (l) 0.651 0.576 0.465Volumetric coefficient of compressor −0.225 −0.225 −0.239Compressor polytropic compression coefficient 2.218 1.897 2.131Electro-mechanical efficiency of compressor(%) 98.41 95.26 93.11Degree of superheat at compressor suction (◦C) 11.1 11.1 11.1Volume of refrigerant in condenser coil (l) 3.00 3.00 2.50Volume of refrigerant in evaporator coil (l) 3.00 3.00 2.50Volume of refrigerant in liquid line (l) 0.20 0.15 0.15Overall heat transfer value of condenser coil (kW K−1) 1.18 1.00 1.00Overall heat transfer value of evaporator coil (kW K−1) 0.88 0.75 0.70Refrigerant charge (kg) 1.38 1.00 1.38Design COP 3.00 3.24 3.42

0.611 0.556 0.472150 138 130130 125 118

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Table 4Total primary consumption without part-load control (kWh).

Office Classroom Shop

Mixing ventilation (MV) 35,248 57,374 67,455Displacement ventilation (DV) 26,221 33,437 44,675

the part-load controller is more significant for mixing ventilationfollowed by stratum ventilation, and the energy saving potentialdrops with the use of the part-load controller.

Fig. 6 shows the year-round-averaged COP for the different casesinvestigated. The percentage improvement in the COP is at least 37%

Auxiliary equipmentCooling tower air volume flowrate (m3/s)Condenser water pump head (kPa)Chilled water pump head (kPa)

ifferent systems vary significantly, the design chilled water sup-ly temperature should also be different. The higher design supplyir temperatures of displacement ventilation and stratum ventila-ion allow higher chilled water temperatures to be adopted. Thismproves the coefficient of performance (COP) of the chiller andffers significant energy saving. To provide the same reasonableifference between the air temperature and chilled water temper-ture at the cooling coil, the corresponding design chilled watereturn temperatures are 13 ◦C for mixing ventilation, 18 ◦C for dis-lacement ventilation and 21 ◦C for stratum ventilation. The designondenser water return temperature is 30 ◦C in all cases. To set aomparable baseline in selecting the chiller parameters, COP of thehillers are assumed to achieve a rated of 3 when the chilled watereturn temperature is 13 ◦C in all cases. The design COP can then beetermined based on the design chilled water return temperatureonly applicable for displacement ventilation and for stratum ven-ilation). The condenser and chilled water flow rates are selectedo that the temperature changes across the chiller are around 5 ◦C.able 3 indicates values of various parameters selected for the air-onditioning systems used in the office under different ventilationethods. Values for classroom and retail shop are selected using

he same principle. Only single chiller is used for each case. Theupply air fan head is taken as 750 Pa with a fan efficiency of 70%n all cases. The cooling tower fan head is taken as 200 Pa with aan efficiency of 65%. All water pump efficiencies are assumed to be0%.

. Methodology of analysis

The TRNSYS simulation software is used for the analysis. NewRNSYS component is developed for the variable-speed chiller. Sys-em simulations are made for one year based on the typical weatherata of Hong Kong [26] using a simulation time step of 3 min. Theotal primary energy consumption and system performance for thearious cases with and without the part-load control are compared.n energy efficiency of 33% is assumed for the electric power plant

n relation to the primary energy input.

. Results and discussion

Table 4 summarizes the year-round-totaled energy consump-ion for the various cases investigated without part-load control.he percentage primary energy saving is at least 25% with the adop-ion of displacement ventilation. This result is consistent with the

ajority of the existing literature. Whereas the percentage primary

Stratum ventilation (SV) 19,609 14,510 25,351Saving from MV to DV 25.61% 41.72% 33.77%Saving from DV to SV 25.22% 56.60% 43.25%Saving from MV to SV 44.37% 74.71% 62.42%

energy saving is 44% with the use of stratum ventilation when com-pared with mixing ventilation. Stratum ventilation offers at least25% energy saving when compared with displacement ventilation.With a higher fresh air ratio like the classroom, the percentageenergy saving can be even up to 74% for stratum ventilation. Thesaving is substantial even taking account additional dehumidifi-cation of the fresh air might be needed. With the functioning ofthe part-load controller, the primary energy consumptions gener-ally decrease due to a better part-load performance except for thestratum ventilation system used in the classroom. An in-depth anal-ysis shows that the energy consumption from the chiller actuallyreduces, but is not sufficient to outweigh the additional consump-tion from the pumps and cooling tower due to a longer runninghour for the chiller. The pumps and cooling tower operate in fullcapacity even if the chiller runs in partial capacity. The benefit of

Fig. 6. Year-round mean COP for different cases.

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ig. 7. Monthly-averaged zone conditions for office concourse without part-loadontrol.

ith the operation of the part-load controller (plc). The COP is theowest for use in the classroom and is the highest for application inhe office.

Fig. 7 indicates the variation of the monthly averaged zone con-itions for the different ventilation methods used in the officeoncourse without the operation of the part-load controller (theart-load control has very little effect on the average zone condi-ions). The variations in the other building types are of the samerend. The temperature change only mildly throughout the yearnder the temperature control, but the relative humidity variesidely which reaches a maximum in the peak-load season. With

he adoption of displacement ventilation or stratum ventilation,he relative humidity increases significantly due to the higher sup-ly air temperatures used. With cooling-only dehumidification, theigher supply air temperature, the more moisture is introduced

nto the building zone. The situation becomes worse for the class-oom and the retail shop cases where the fresh air ratio is high. Theeak zone relative humidity in the classroom can go up to nearly5% with the use of stratum ventilation. A high relative humidity

mpairs the indoor air quality besides the thermal comfort. To solvehe problem, an auxiliary system is needed to work in parallel withhe chiller to handle the zone latent load. A proper dehumidifica-ion process should be in place to suit a specific ventilation method.here is a general trend to separate dehumidification from theooling process since 1990s [27]. A desiccant dehumidification sys-em, which utilizes low-grade heat for regeneration, is one possibleption. The use of solar energy offers extra benefit in terms of theverall system energy efficiency. Further works can be proceeded toxplore the feasibility. Even if the energy used for dehumidifications considered, the energy saving potential of stratum ventilation istill substantial. On the other hand, the higher the zone tempera-ure, the higher moisture content in the air for a specific relativeumidity. This reduces the dehumidification load for stratum ven-ilation.

. Conclusions

The energy simulation software TRNSYS is used for computa-ion of the space cooling load and system energy consumption.ear-round energy consumptions of three systems of stratumentilation, displacement ventilation and mixing ventilation areompared for typical configurations of an office, a classroom andretail shop in Hong Kong. The energy saving potential of stratumentilation is derived mainly from three factors: the reduction in

entilation and transmission loads and increased COP of chillers.ecause of the higher supply air temperature of stratum ventila-ion, the energy consumed for the treatment of fresh air is mucheduced. Similarly, elevated room temperature leads to smalleremperature difference between the indoors and the outdoors, and

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ngs 43 (2011) 130–136 135

therefore lower transmission. The increased COP reduces chillerpower consumption associated with stratum ventilation. Throughthe year-round energy simulation of the typical office, classroomand retail shop in Hong Kong, stratum ventilation has shown tobe able to provide substantial energy saving when compared withconventional ventilation methods. The figure is at least 25% and44% when compared with displacement ventilation and mixingventilation, respectively.

Acknowledgement

The work described in this paper is supported by a StrategicResearch Grant of the City University of Hong Kong, Hong KongSpecial Administrative Region, China (Project No. 7008038).

References

[1] Electrical and Mechanical Services Department (EMSD), Government of HongKong S.A.R., Energy Conservation Charter 2006—Suitable Room Temperature,http://www.energyland.emsd.gov.hk/blue sky/eng/index.htm.

[2] National Development and Reform Commission (NDRC), the State Coun-cil of China, Citizen’s Pact for Deed of Thrift, http://www.ndrc.gov.cn/gzdt/t20051223 54462.htm.

[3] NOW news, Taiwan, Campaign in office of president for saving energy andreducing carbon emissions, http://www.nownews.com/2008/05/24/10844 -227 9584.htm.

[4] Ministry of Knowledge and Economy, the Republic of Korea.The standard of energy management, December 29, 2008.http://www.mke.go.kr/info/law/gosiView.jsp? (Item No. 2008-219).

[5] Japanese Ministry of the Environment (MoE), National Campaign to Fight GlobalWarming, http://www.env.go.jp/en/press/2005/0428b.html.

[6] International Facility Management Association. Temperature wars—savings vs.comfort, July 2009. http://www.ifma.org.

[7] Z. Lin, T.T. Chow, C.F. Tsang, Stratum ventilation? A conceptual introduction,in: Proceedings of the 10th International Conference on Indoor Air Quality andClimate, Beijing, China, September 4–9, 2005, pp. 3260–3264.

[8] Z. Lin, T.T. Chow, C.F. Tsang, K.F. Fang, L.S. Chan, Stratum ventilation—a potentialsolution to elevated indoor temperature, Building and Environment 44 (11)(2009) 2256–2269.

[9] Z. Lin, L. Tian, T. Yao, Q. Wang, T.T. Chow, Experimental and numerical study ofroom airflow under stratum ventilation, Building and Environment, in press,doi:10.1016/j.buildenv.2010.07.018.

10] L. Tian, Z. Lin, J. Liu, Q.W. Wang, Numerical study of indoor air quality andthermal comfort under stratum ventilation, Progress in Computational FluidDynamics 8 (7/8) (2008) 541–548.

11] L. Tian, Z. Lin, Q.W. Wang, J. Liu, Comparison of gaseous contaminant diffusionunder stratum ventilation and under displacement ventilation, Building andEnvironment 45 (9) (2010) 2035–2046.

12] L. Tian, Z. Lin, Q.W. Wang, J. Liu, Numerical investigation of indoor aerosol par-ticle dispersion under stratum ventilation and under displacement ventilation,Indoor and Built Environment 18 (4) (2009) 360–375.

13] D.P. Wyon, The effects of indoor air quality on performance and productivity,Indoor Air 4 (Suppl. 7) (2004) 92–101.

14] Y.-P. Ke, S.A. Mumma, D. Stanke, Simulation results and analysis of eight venti-lation control strategies in VAV systems, ASHRAE Transactions 103 (2) (1997)381.

15] A.M. Zhivov, A.A. Rymkevich, Comparison of heating and cooling energy con-sumption by HVAC system with mixing and displacement air distribution fora restaurant dining area in different climates, ASHRAE Transactions 104 (2)(1998) 473–484.

16] G.A. Florides, S.A. Tassou, S.A. Kalogirou, L.C. Wrobel, Measures used to lowerbuilding energy consumption and their cost effectiveness, Applied Energy 73(3–4) (2002) 299–328.

17] J. Niu, L. Zhang, H. Zuo, Analysis of energy and humidity performance of a systemcombining chilled ceiling with desiccant cooling, ASHRAE Transactions 108 (2)(2002) 195.

18] H. Xu, J. Niu, Numerical procedure for predicting annual energy consump-tion of the under-floor air distribution system, Energy and Buildings 38 (2006)641–647.

19] F. Karlsson, P. Rohdin, M.-L. Persson, Measured and predicted energy demandof a low energy building: important aspects when using building energy sim-ulation, Building Services Engineering Research and Technology 28 (3) (2007)223–235.

20] C.K. Lee, Dynamic performance of ground-source heat pumps fitted withfrequency inverters for part-load control, Applied Energy (2010) 029,doi:10.1016/j.apenergy.2010.04.

21] H. Jin, J.D. Spitler, A parameter estimation based model of water-to-water heatpumps for use in energy calculation programs, ASHRAE Transactions 108 (1)(2002) 3–17.

22] P. Domanski, D. Didion, Mathematical model of an air-to-air heat pumpequipped with a capillary tube, International Journal of Refrigeration 7 (4)(1984) 249–255.

1 Buildi

[

[

[

[

36 Z. Lin et al. / Energy and

23] H. Skistad, Displacement Ventilation, Research Studies Press Ltd., Taunton,

1994.

24] L. Tian, Experimental and numerical study of indoor environment under stra-tum ventilation, Doctoral Thesis, Xi’an Jiaotong University, 2009.

25] R. Kosonen, Displacement ventilation for room air moisture control in hot andhumid climate air distribution techniques (Air distribution techniques 5), in:HVAC Handbook—Displacement Ventilation in Hot and Humid Climate, 2010.

[

ngs 43 (2011) 130–136

26] A.L.S. Chan, T.T. Chow, S.K.F. Fong, J.Z. Lin, Generation of a typical meteorological

year for Hong Kong, Energy Conversion and Management 47 (1) (2006) 87–96.

27] A.A. Pesaran, A review of desiccant dehumidification technology, in: Proceed-ings of EPRI’s Electric Dehumidification: Energy Efficient Humidity Control forCommercial and Institutional Buildings Conference, New Orleans, Louisiana,June 2–3, 1993.