Energy performance of next-generation dedicated outdoor air cooling systems in low-energy building operations

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Energy Performance of Next-Generation Dedicated Outdoor Air Cooling Systems in Low-Energy Building Operations Tianyi Chen , Leslie Norford PII: DOI: Reference:

S0378-7788(19)31875-4 https://doi.org/10.1016/j.enbuild.2019.109677 ENB 109677

To appear in:

Energy & Buildings

Received date: Revised date: Accepted date:

14 June 2019 19 October 2019 4 December 2019

Please cite this article as: Tianyi Chen , Leslie Norford , Energy Performance of Next-Generation Dedicated Outdoor Air Cooling Systems in Low-Energy Building Operations, Energy & Buildings (2019), doi: https://doi.org/10.1016/j.enbuild.2019.109677

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Highlights



Development of thermodynamic models of desiccant and membrane cooling systems appropriate for seasonal, climate-specific performance assessment

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Two designs for membrane systems identify opportunities to reduce work associated with vacuum pump needed to maintain a difference in water-vapor partial pressure across the membrane.

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COP distribution curves quantify the superior energy performance in latent heat removal provided by desiccant and membrane systems

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Energy recovery and renewable energy systems contribute enhanced system energy efficiency, notably the use of ground exchange when permitted by climate-specific deep-ground temperatures

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Energy Performance of Next-Generation Dedicated Outdoor Air Cooling Systems in Low-Energy Building Operations Tianyi Chen a*, Leslie Norfordb a: Department of Mechanical Engineering, Massachusetts Institute of Technology b: Department of Architecture, Massachusetts Institute of Technology *Corresponding author; Email: [email protected]

Abstract A conventional chiller system based on the vapor-compression cycle, widely used in dehumidification and cooling, has its limitations in energy performance and environmental concerns. Previous studies have mainly focused on analysing one specific type of alternative cooling technology in a certain climate or for a specific building. In this study, we apply thermodynamic models to systematically evaluate the energy efficiencies of next-generation dedicated outdoor air cooling systems, namely those with desiccants and membranes, and compare them against the industry benchmark, revealing their prospects and limitations at different time scales and locations. The innovative systems we proposed are coupled with energy recovery and renewable energy systems to further enhance the system energy efficiencies. A sensitivity analysis on equipment performances further informs how to choose equipment parameter values to boost the system energy performances. Results reveal the principles of how to systematically organize cooling equipment to achieve potential energy savings from a thermodynamic point of view, and highlight 100% increase in working-hour average system COP for desiccant and membrane cooling systems exampled in a low-energy office building at Boston. Parametric studies in different climates and working conditions further substantiates the superior system energy performances of next-generation cooling technologies in latent heat removal.

Keywords Dedicated outdoor air cooling systems, energy efficiency, low-energy buildings

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1. Introduction For over a century, vapor compression systems (VCS), featured in air conditioners and chillers, have been commonly used to supply conditioned air for indoor thermal comfort. However, the energy inefficiencies and environmental concerns about refrigerants in the VCS system have led to a wide search for alternatives. Chua et al. [4] reviewed the latest innovations on HVAC systems to improve energy efficiency. The U.S. Department of Energy (DOE) published a study that compares the technical energy saving potentials of 17 non-VCS alternatives, among which desiccant- and membrane-based technologies are ranked as highly promising [5]. Goetzler et al. [6] further highlighted the cooling saving potentials of the two technologies, which in some cases have over 50% higher efficiencies than the VCS. To quantify the energy savings of different HVAC technologies, thermodynamic analyses are necessary to understand the energy conversion among each component in the systems. Labban et al. [7] conducted a comparative study of next-generation HVAC cooling systems to thermodynamically understand the prospects and limitations of desiccant and membrane cooling systems with respect to the VCS. Bui et al. [8] analyzed the energy performance of a vacuum-based membrane dehumidification system, identifying an achievable order-of-magnitude improvement in efficiency relative to the VCS. While significant progress has been made to understand the energy efficiencies of different HVAC systems, previous studies have either focused on one specific technology or compared performance under idealized, static building operations with no inclusion of dynamic indoor cooling loads. There remain questions about how these nextgeneration cooling systems, especially desiccant- and membrane-based cooling systems, perform in realistic building operations and how climates can potentially change the energy performance of different systems that integrate renewable energy systems and energy recuperation. Our work evaluates the energy saving potentials of next-generation dehumidification and cooling systems, namely those with desiccants and membranes, compared with the industrial benchmark, the baseline chiller system with vapor-compression cycle. We focus exclusively on dedicated outdoor air systems that condition outdoor air brought into a building to meet air quality standards. We construct innovative nextgeneration DOAS cooling systems coupled with energy recovery and renewable energy systems, and conduct thermodynamic models to study the system energy performance operated in dynamic working conditions. Impacts of equipment performance, building operations and climates are assessed, and results highlight the improved system energy performances for the next-generation cooling systems in latent heat removal. The originality of our work includes: 1) development of an evaluation platform to systematically and fairly compare the energy performances of different DOAS cooling systems, both in varying times scales and locations; 2) design and modeling of innovative next-generation cooling systems with energy recovery and renewable energy resources to enhance energy efficiencies; 3) a sensitive analysis to reveal how to optimize the design of equipment for different systems to achieve boosted system energy performance; 4) compelling case studies exemplified in Boston, and parametric studies further extended to various climates and working conditions, for next-generation dehumidification and cooling technologies in latent heat removal.

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2. Nomenclature Symbols ̇ ̇

Abbreviation BCS DCS DOAS DW ECS ERW GHE HCS HW MCCS

Mass flow rate of moist air, kg/s Mole flow rate of the water vapor across the membrane, mol/s Flat-sheet membrane plan area, m2 Thermal capacity of the moist air, kJ/K Thermal capacity of water, kJ/K Average feed-side water vapor partial pressure along the airflow direction, Pa Vacuum pressure at the entrance of the vacuum pump, Pa Total cooling load of the DOAS, kW Process air temperature at the inlet of the desiccant wheel, K Process air temperature at the outlet of the desiccant wheel, K Chilled water temperature leaving the evaporator, ℃ Heating work (from heat pump), kW Sensible cooling work (from electric chiller or ground heat exchanger), kW Compression work by the liquid compressor, kW Condensation cooling work, kW Fan power, kW Vacuum pump work, kW Number of flat-sheet membrane stacks Water permeance of the membrane, mol/(s·m2·Pa) Effectiveness of heat exchanger Latent heat recovery ratio Sensible heat recovery ratio Isentropic efficiency of the vacuum pump Non-dimensional outlet temperature of the desiccant wheel Specific enthalpy of the moist air, kJ/kg Specific entropy of the water vapor, kJ/(kg K) Humidity ratio of the air, kg/kg (dry air)

Baseline chiller system Desiccant cooling system Dedicated outdoor air system Desiccant wheel Baseline chiller system coupled with enthalpy recovery wheel Enthalpy recovery wheel Ground heat exchanger Baseline chiller system coupled with heat recovery wheel Heat recovery wheel Membrane cooling system with condenser

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MCMS

Membrane cooling system with return membrane

3. DOAS Cooling System Design This section describes different approaches to dehumidifying and cooling the indoor environment: the conventional vapor-compression cycle, and two promising next-generation dehumidification technologies, desiccants and membranes. Six system designs are presented with energy recovery and renewable energy systems to achieve enhanced energy performances. The DOAS, shown in Figure 1, supplies cool and dry fresh air with constant volumetric flow rate to satisfy indoor air quality and thermal comfort requirements, and completely removes the indoor latent heat. A separate indoor sensible cooling system, such as chilled beam or radiant ceiling, can meet the additional indoor sensible cooling load, which is not included in the following analysis.

Figure 1. Schematic of a DOAS. The baseline chiller system (BCS), shown in Figure 2(a), implements the most commonly used dehumidification and cooling technology, the vapor-compression cycle, which we designate as the industrial benchmark to compare with innovative systems. The BCS normally provides chilled water under 10 ℃ to sensibly cool the outdoor air to the dew point temperature at State 2, and further condense the water vapor out for latent cooling. The overcooled dehumidified air at State 3 will be conditioned to the room temperature with waste heat. The chilled water is supplied by the vapor-compression cycle where cooling work is paid by compressing the refrigerant from evaporator to condenser. The heat absorbed in the coolant is rejected to the condenser where the coolant water carries heat away and is evaporatively cooled by the cooling power. To remove the latent heat from the outdoor air, the BCS requires substantial cooling work to generate low-temperature coolant in order to overcool the air to the dew-point temperature. Figure 2(b) - (c) depict two variations of the BCS with integrated heat /enthalpy recovery wheel to improve the energy efficiency. The heat/enthalpy recovery wheel, a rotary heat exchanger, consists of a circular honeycomb matrix of heat-absorbing material (and water absorbent for enthalpy recovery), slowly rotated to transfer sensible/sensible + latent heat from hot and humid outdoor air to cool and dry indoor exhaust air, before entering the BCS (framed in red box). Since the heat is transferred without loss in the exchange medium, the efficiencies of the wheels are usually higher than other forms of heat exchangers [9]. In the proposed systems, heat/enthalpy recovery pre-conditions the outdoor air with minimal energy input, and will potentially reduce the sensible (and latent) cooling loads on the BCS,

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which improves its performance with warmer coolant supply. In the working conditions when outdoor air is cooler (or less humid) than the indoor air, the energy recovery system will be bypassed to avoid accidental heating or humidification. A desiccant-based dehumidification and cooling system (shown in Figure 2(d)), separates latent and sensible cooling by utilizing advanced water-absorbing materials. The latent cooling is processed in the desiccant wheel (DW), a honeycomb device with multiple internal channels coated with desiccant materials (typically silicone gel or zeolite) and rotating constantly to alternate periodically between the dehumidification and cooling process. It is favored in household and office use due to its compactness, continuous operation and lower susceptibility to corrosion compared with liquid desiccants [14]. Due to the adsorption heat incident on the surface of the channel, the dehumidified air has high temperature which requires additional sensible cooling to reduce its temperature to the room condition. At the same time, a continuously running desiccant wheel also needs hot regeneration air to remove the water vapor from the desiccant material. For the DCS proposed in Figure 2(d), a heat recovery wheel (HW) is used to provide sensible cooling with the cool indoor air, and supplies heated exhaust air to regenerate the DW. A cooling coil and a heat pump are installed downstream in the process air and the exhaust air respectively to provide on-demand conditioning. For climates with seasonal temperature variation or strong diurnal patterns, the ground is a good thermal storage medium, where it can provide cooling and store heat during summer or the peak hours of the day, and then flush out heat during winter or nighttime. Therefore, we propose to use the renewable energy system, the ground heat exchanger (GHE) to provide sensible cooling with high energy efficiency in those climates. For tropical climates with high ground temperature, a GHE cannot work effectively and the chiller system presented before can be used as the alternative. The efficiency of the BCS can be greatly enhanced when it serves sensible cooling only with warm coolant supply. Membrane-based dehumidification and cooling systems (shown in Figure 2(e) - (f)), separates latent and sensible cooling by utilizing advanced selectively permeable materials. The latent cooling is processed in the membrane module, which has a hydrophilic semi-permeable material coated on a substrate and maintains a water vapor pressure gradient across the two sides [23][24][25]. Unlike the desiccant that dehumidifies and heats the process air, membrane dehumidification is advantageous due to its isothermal separation of water vapor from humid air, when hydraulic losses across the membrane are neglected. A vacuum pump is installed on the permeate side of the membrane, providing the required pressure gradient to separate the excessive water vapor in the outdoor air. To prevent condensation in the vacuum pump associated with directly pressurizing the water vapor to the atmosphere pressure, we propose two system designs: one is to condense water vapor to liquid water at an intermediate pressure [28] (shown as the MCCS in Figure 2(e)); and the other is to force the water vapor through the bottom membrane (shown as the MCMS in Figure 2(f)) at low pressure and mix it with the exhaust air for rejection. The MCCS uses an enthalpy recovery wheel (ERW), upstream of the membrane, to pre-cool and pre-dehumidify the outdoor air with cool and dry exhaust air. The energy-efficient pre-conditioning of the outdoor air reduces the humidity removal rate at the membrane, and hence reduces the overall work paid for the system. Like the DCS, a cooling coil provides necessary sensible cooling with the chilled water supplied by the GHE or the BCS as described above, at a much lower load compared with the BCS and its variations. A pump is installed downstream of the condenser to pressurize the liquid condensate to the ambient environment, whilst reusing the collected water for sewage. Inspired by Claridge and Culp [34][35], the substitution of

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an exhaust-air membrane for the condenser and second-stage compressor in the MCMS exchanges the thermal work input for purely mechanical work by the vacuum pump. It achieves dehumidification and cooling without any form of condensation. For the MCMS, the ERW is not implemented upstream of the membrane module (as referenced to the process air stream), because the humidity recovered in the exhaust air will increase the water vapor partial pressure at State 7 and therefore increase the vacuum pump work to maintain the same pressure gradient. A HW is used instead downstream of the membrane to recuperate the waste heat from the hot dehumidified air. To reduce the risk of fouling the lower membrane by condensation, the MCMS uses an on-demand heat pump to raise the temperature of the exhaust air and its capacity to absorb water vapor, when the water vapor removal rate is higher than the allowable amount at State 8 before saturation.

1 Outdoor air

4 2

Exhaust air

3 Supply air

(a)

1

2

Outdoor air Exhaust air 6 5 3

4 Supply air

(b)

1

2

5 3

(c)

7

Outdoor air

6 Exhaust air

4 Supply air

8 1

Outdoor air

Exhaust air

6

5

7

4 Supply air

3

2

(d)

1

Outdoor air

6

2 Exhaust air 5 4 3 Supply air

(e)

8 1

Outdoor air

Exhaust air

5 4 3 Supply air

6 (7) 2

(f) Figure 2. Schematic of the six proposed DOAS cooling systems. (a) to (f) describe the system designs and the psychrometric chart of air conditioning process for the baseline chiller system (BCS), BCS with heat recovery wheel (HCS), BCS with enthalpy recovery wheel (ECS), desiccant cooling system (DCS), membrane cooling system with condenser (MCCS), membrane cooling system with membrane (MCMS), respectively.

4. Thermodynamic Modeling of DOAS The energy performance of the DOAS is evaluated by the Coefficient of Performance (COP), defined as the ratio of the overall cooling load incident on the DOAS system to the total cooling work paid for the system. The total cooling load of the DOAS is the product of the mass flow rate of air and the enthalpy difference between the outdoor air and the supply air (with room temperature and lower humidity ratios than the indoor environment for indoor latent cooling). For a fair comparison among different systems, the overall system cooling load is identical. The total cooling work consists of fixed fan power (including supply fans, exhaust fans, and fan-powered terminal units associated with the DOAS systems

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to drive the airflow in the duct with pressure drop of 1.5 kPa [9]) and other forms of energy input depending on the system design, e.g. compressor/pump work, etc. Equations Error! Reference source not found. - (7) show the system-specific expressions for the COP, where the fan power includes both fixed fan power and the equipment fan power to overcome the pressure drop across the equipment due to the friction loss 1. (1) (2) (3) (4) Note that the BCS and its variations share the same expressions for the system COP, where the sensible cooling work of the HCS and ECS is no greater than the BCS according to the thermodynamic states of the pre-conditioned air characterized by and , evaluated as follows. (5) (6) where the thermodynamic state index refers to Figure 2(b) - (c).

4.1 Electric chiller model The sensible cooling work by the electric chiller comes from the compressor work to pump the refrigerant from the evaporator to the condenser chamber. In order to achieve an analytical estimation in practice, we chose to use the electric chiller model presented in the EnergyPlus Engineering Reference [10] with details described in the appendix. This model requires the two inputs of chilled water temperature and entering condenser temperature. Given the value of heat exchanger effectiveness, we can then determine the required chilled water temperature (in Figure 2(a)) to dehumidify the air to the desired humidity level, as noted in Equation (7): ( ) (7) ) { }( Because the discharge air is saturated, is the dew point temperature at the desired humidity ratio. The cooling tower can cool the condenser water to the wet-bulb temperature of the outdoor air, and thus the entering condenser temperature is approximated as the web-bulb temperature plus an approach temperature of 4 °C [11][12][13].

1

The pressure drop across the equipment depends on mass flow of the air, equipment size and detailed configurations. We refer to the technical specifics of the product by the Desiccant Rotors International (DRI) company (http://www.drirotors.com/prod_detail.php?prod_id=1) to estimate the friction loss for the DW and energy recovery devices.

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4.2 Desiccant dehumidification model Performing an energy analysis of the desiccant wheel requires a detailed description of the heat and mass transfer process for adsorption/desorption on the desiccant material. Ge et al. [15] reviewed different approaches to solving the equations numerically or analytically with simplified assumptions. While numerical simulations can present detailed physical descriptions of heat and mass transfer processes in the desiccant wheel, analytical simplifications provide useful insights to understand the process and reduce the complexity of the problem. The literature shows two analytical approaches. One is based on empirical studies of desiccant wheel performance, either defining the effectiveness of thermal potential as an analogy to a heat exchanger model [7][16] or using polynomial fits [17] to calculate the outlet air conditions. Both have limitations in dealing with a wide range of outdoor air conditions, especially when the DOAS is required to remove large amounts of indoor latent loads. The other analytical approach, which we adopt in this paper, is based on further assumptions of linear and steady air temperature and humidity ratio profiles in the desiccant channel in typical working conditions ( s), which are validated against the experimental studies with time-average absolute error less than 3% [18]. Explicit analytical solutions can be achieved by introducing non-dimensional variables and a set of dimensionless parameters and coefficients for the desiccant wheel, as is shown in Table A-2. By calculating physical coefficients for the desiccant and integrating over the whole processing time the mean average outlet air temperature can be expressed as follows: ( ) ( ) ̅̅̅̅̅̅̅ ∫ ⁄ ⁄ (8) ( )) ( )) ( ( ⁄ where

⁄

is the non-dimensional outlet temperature of the desiccant wheel,

describing how far the actual dehumidification process deviates from the idealized isenthalpic process measured by the temperature gradient of the outlet and inlet air temperature. From the expressions above, we can calculate the outlet process air temperature at different working conditions by fixing ̅̅̅̅̅̅̅ . Constant ̅̅̅̅̅̅̅ can be achieved by regulating rotational speed of the desiccant wheel according to hourly latent load removal rates, since the rotational speed directly affects the processing time of the desiccant wheel and thus influences thermal and sorption capacity of the wheel. Besides the fan power needed to overcome the friction loss across the DW, another source of work input of the desiccant dehumidifier is the regeneration work by the heat pump. The heating work in Equation (9) is computed with a nominal COP of 4.8 for the top-rated heat pump provided in ASHRAE Standard 90.1-2016 [9] and a high-temperature heat pump provided in the review paper [22]. ) ̇ ( (9) where the regeneration temperature at State 7 (in Figure 2(d)) can be backtracked to the intersection of the isohume line starting from the ideal point and the constant humidity ratio line at [18]. Notice that the humidity ratio of the exhaust air remains the same until it takes in water vapor in the regeneration process, i.e. . Given the sensible heat recovery of the HW, the inlet air temperature at the cooling

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coil is then determined by Equation (5). Drawing an energy balance on the HW gives the exhaust air temperature at State 6. (10) Then, the heating work can be calculated directly from Equation (9). For some situations when outdoor air is quite dry and the overall latent cooling load of the DOAS system approaches zero, there is no need to dehumidify the outdoor air. Then, the outdoor air will bypass the desiccant wheel and directly flow into heat recovery wheel and the succeeding cooling coil for sensible cooling only.

4.3 Membrane Dehumidification Model In the literature, there are two forms of membrane module used for dehumidification and cooling: flatsheet membrane [24][26] and hollow-fiber membrane [23][25]. The flat-sheet membrane is commonly used for its simpler structure, smaller pressure loss and lower CAPEX [25], and thus we adopt this form in the following analysis. Membrane dehumidification is a mass transfer process with the driving force from the transmembrane pressure gradient, influenced by the permeability and selectivity of the membrane. Recent studies have reported achieving a very high selectivity of water vapor to air molecules (almost 1,000) for a polymer membrane [26], and thus we assume that the membrane can perfectly separate water vapor from moist air2. Compared with the simplified model with constant permeate-side pressure and infinite membrane area for humidity removal [7], we consider the decrease in water vapor partial pressure across the feed side, and estimate the uniform vacuum pressure as follows. ( ) ̇ (11) According to mass conservation of water vapor across the membrane given steady-state operations and uniform permeate pressure, the mole fraction of water vapor and the water vapor partial pressure experience an exponential decrease along the flow direction [27][28], assuming constant working pressure of the feed-side air3. The integrated average of the feed-side water vapor partial pressure can thus be expressed as, ∫

(

)

(

)⁄(

)

(12)

where and refer to the inlet and outlet water vapor partial pressure (unit: Pa), is the flowdirection length of the membrane, respectively. The airside pressure drop along the length of the membrane sheet due to friction can be calculated from the Darcy-Weisbach equation [29][30]. When the overall latent cooling load of the DOAS system approaches zero, the outdoor air will also bypass the membrane module and be sensibly cooled by the cooling coil. Given the isentropic efficiency and the exit pressure of the vacuum pump, the thermodynamic state of water vapor at State 8 (in Figure 2(e)) is then determined as follows [7]. (13) ) ̇ ( (14) where is determined by . At the intermediate pressure of 10~100 mbar, the condensation temperature of water vapor is close to the ambient temperature, and thus it will be energy efficient to 2

Further simulation results prove that the total separation work with air permeation only increases by 2% with the water vapor to air selectivity of 1000, which validates the assumption here. 3 The assumption is valid because the pressure drop across the membrane is much smaller than the ambient pressure.

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provide cooled water from the ground heat exchanger used in the sensible cooling or from a top-rated water-cooled chiller. The compression work for the liquid water is the product of the volumetric flow rate of liquid water and the pressure difference between State 9 and 10, which is usually negligible. For the MCMS, the exit pressure of the vacuum pressure is determined by the mass flow balance of the separated water vapor in the system; for the same total membrane area for the two membranes, the pressure gradient across the membrane should be the same. (15) where and refer to the integrated average of the feed-side water vapor partial pressure along the airflow direction, calculated as Equation (12). Given the water vapor removal rate, the water vapor pressure at State 9 and 10 (in Figure 2(f)) can be then determined. The vacuum pump work is calculated similar to Equation (13) - (14) with state numbers replaced by 10 and 9, respectively.

5. Results and Discussion To evaluate the energy performance of DOAS cooling systems in low-energy building operations, we first run the building energy model with the proposed low-energy building prototypes and obtain the timedependent internal cooling loads as well as the indoor and outdoor air conditions to quantify the overall cooling loads on the DOAS systems. Then, we perform a thermodynamic analysis to compare the energy performance of next generation DOAS cooling systems with the conventional chiller system throughout the cooling season. As a case study, the proposed systems are evaluated during annual operations in a low-energy office building prototype in Boston, USA. We further address the impact of equipment sizing and performances on the overall system energy performances through the sensitivity analysis. We extend our discussions to a combination of different climates and internal cooling loads, while uncovering limitations in applicability and feasibility of the DOAS alternatives.

5.1 Low-Energy Building Performance First, we construct building envelopes and run energy simulation to identify the amount of indoor generated latent cooling loads to be further removed. Low-energy buildings with tight envelopes, energyefficient equipment and energy recovery technologies have gained an increasing focus, and zero net energy (ZNE) buildings, both certificated and emerging ones, are widely constructed in the U.S. and Europe [36][37][38][39]. Thus, we apply our analysis in the context of low-energy building operations. Table 1 lists the low-energy office building design that determines indoor latent cooling load generation. Complete design specifics for a low-energy office building prototype in mild and humid climates based on ASHRAE Standard [9] are presented in the appendix Table D. To evaluate the building energy consumption with design specifics, we simplify each floor of the building as a single zone, and run a single zone simulation with EnergyPlus 4 [40][41]. Table 1. Low-energy building prototype and its energy performances. 4

Since there is negligible heat flow with very close indoor temperatures between floors, we can approximate each floor as a thermally isolated zone, except the top and the ground floor.

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Category Value (Unit) Total floor area 1750 (m2) Occupant density 0.1 (p/m2) Mechanical ventilation rate 0.55 (L/s/m2) Infiltration rate 0.6 (/hr) Total cooling load 17.1 (kWh/m2/yr) Total EUI 54.9 (kWh/m2/yr) Note: We assume COP for heating is 3.0, COP for cooling is 3.0 [7]. We compare our simulation results with zero net energy emerging buildings in Massachusetts, USA. It is reported that the Woods Hole Research Center at Falmouth, MA, achieves the site EUI of 50 kWh/m 2/yr, and that Massachusetts Division of Fisheries and Wildlife – Field Headquarters Building achieves the site EUI of 68 kWh/m2/yr [38][39]. Our low-energy office prototype building achieves the site EUI of 54.9 kWh/m2/yr, which validates its use in evaluating system performance in high-efficiency buildings.

5.2 Case Study We conduct a case study to compare real-time energy performances for the three DOAS cooling systems in a low-energy office prototype building in Boston. We first obtain outdoor air conditions from the weather profile and indoor air conditions and indoor latent loads from building energy simulation by EnergyPlus, and then feed them into thermodynamic models of the six proposed systems developed in Modelica [42][43] to simulate the energy performance. The annual average outdoor temperature at Boston is 10.7 °C, lower than the required cooled water supply temperature in the desiccant and membrane cooling systems, and thus the ground heat exchanger can be used for additional sensible cooling in these systems. Figure 3 summarizes the working conditions of the DOAS cooling systems, where (a) and (b) show the frequency and intensity of the systems, respectively. The DOAS cooling systems only work from May to September in Boston, while the total load in August is the highest of the year, peaking around 1.5 kWh/m2. We conclude that the cooling season at Boston is June to August, during which the cooling load of the systems is over 90% of the annual total.

(a)

(b)

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Figure 3. Working conditions of the DOAS cooling systems. (a) depicts the monthly working hours (frequency) of the DOAS systems, while (b) depicts the monthly working loads (intensity) of the systems. Figure 4 shows the COP of different DOAS cooling systems at different time scales. Note that COP = 0 when the DOAS cooling system is off. We can clearly see from the monthly and diurnal COP patterns that the three proposed next-generation DOAS cooling systems (DCS, MCCS and MCMS) outperform the BCS and its variations at Boston during the most of the summer time with the highest COP achieving 17.8 for the DCS. The spikes of the COP curve indicate the inconsistent need of active cooling in Boston during the summer, as well as the peak load incident at the start of the working hours due to the closure of HVAC systems at night. Comparing the six systems in a typical workday in August, we can see a strong diurnal pattern in the energy efficiencies due to the diurnal changes of outdoor air conditions and occupant density. The efficiency spikes happen in the late morning and afternoon, when outdoor air temperature increases and the indoor latent loads rise to the peak with the maximum indoor activity level. Low cooling loads always lead to low energy efficiency owing to the fixed fan power requirements for fresh air intake. Of the six proposed systems, the DCS and the MCCS take the lead in low energy consumption with COP usually above 10 during the working hours, followed by the MCMS with COP around 8, the ECS with COP around 6, and the BCS and the HCS with lowest efficiencies. COP of DOAS Cooling Systems in a Cooling Day at Boston

COP of DOAS Cooling Systems in August at Boston 20

18

18

16

16

14 12

COP

COP

14 BCS

12

HCS 10

ECS

8

BCS 10

HCS

ECS

8

DCS

DCS

MCCS

6

MCMS 4

6

MCCS

MCMS

4

2

2 0

0 0

150

300

450

600

750

0

Time (Hour)

4

8

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24

Time (Hour)

(a) (b) Figure 4. Comparison of the time-variant energy performance of the six DOAS cooling systems. (a) refers to the COP of different DOAS systems operating in August at Boston; (b) refers to the COP in August, 4th. Figure 5 (a) compares the COP distribution for different DOAS cooling systems during their working hours in the year. The peaks in the density plot, namely the mode COP of the density distribution, best represent the energy saving potentials of the proposed desiccant and membrane cooling systems compared to the baseline chiller system and its variations, with a remarkable three-fold increase in energy efficiencies for the DCS and the MCCS (peak-density COP of approximately nine, compared to three for the base system) and over 100% increase for the MCMS. The peaks always occur at high cooling loads, which proves the superior system energy performance of the next-generation DOAS cooling systems in removing latent heat. Figure 5 (b) compares the cooling work distribution for the six DOAS cooling systems. The desiccant and membrane cooling systems have much smaller variances, concentrated close

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to 2 kW, while the BCS and its variations spread out to 7 kW during operations, marking their instability in power requirement.

(a) (b) Figure 5. Comparison of the distribution curves of the proposed DOAS cooling systems at working hours in Boston. (a) depicts the COP distribution, while (b) depicts the cooling work distribution, calculated as the ratio of the total DOAS cooling load to the COP. Comparisons of the working-hour average COP of the proposed DOAS cooling systems further validate the findings stated above, as shown in Table 2. The working-hour average COP is defined as the ratio of the total annual cooling load to the total annual work input, provided as a time-average metric to fairly compare different system performances. The proposed next-generation DOAS cooling systems, including DCS, MCCS and MCMS, have nearly twice the energy efficiency of the BCS and its variations, proving their great potential in reducing cooling energy consumption in Boston. The enthalpy recovery wheel combined with the BCS is more energy-efficient than the heat recovery wheel due to its coupled removal of sensible and latent cooling loads; however, no significant improvement results from the mild outdoor air temperature when wheels are bypassed. The MCMS is less competitive than the MCCS and the DCS because the humid outdoor air at cooling season increases the vacuum pump work to reject the large amounts of the water vapor to the returning membrane. Table 2. Summary of energy performances of different DOAS cooling systems in Boston. Total Cooling Loads (kWh/m2/yr) 3.09

Total Working Hours (h) 858

Average DOAS COP in Working Hours BCS

HCS

ECS

DCS

MCCS

MCMS

3.73

3.74

3.93

7.23

7.99

6.58

5.3 Sensitivity Analysis To quantify the impact of equipment sizing and performance on the overall system energy performance, as characterized by system COP, we perform a parametric study for a cooling system in Boston. Results are summarized in Figure 6.

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(a)

(b)

(c)

(d)

(e) (f) Figure 6. Effects of equipment sizing and performances on the system energy performance for different DOAS cooling systems, evaluated by working-hour average COP.

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Figure 6(a) shows the energy performance of the DCS with different desiccant wheel configurations characterized by the dehumidified air temperature. As the outlet dehumidified air temperature rises, the working-hour average COP of the DCS initially increases, climbs to a peak, and gradually decreases at high air temperature. The result reflects the tradeoff between the regeneration work and the sensible cooling work. Low dehumidified air temperature will reduce the amount of heat recovered by the heat exchanger, and thus the system performance is limited by the high regeneration work from the heat pump. High dehumidified air produces recoverable heat that reduces regeneration work; however, the increase in sensible cooling work to cool the process air to the room temperature will lower the system COP. In addition, the maximum system COP becomes higher with increasing sensible heat recovery ratios, accompanied with higher dehumidified air temperature. It reflects the high dependence of the system performance on the heat recovery ratio of the HW, which measures how much regeneration work and sensible cooling work can be offset by exchanging heat with the cool exhaust air. When the sensible heat recovery ratio is smaller than 0.70, high dehumidified air temperature will significantly increase both regeneration work and sensible cooling work, and result in energy-intensive operations. This study shows that moderate dehumidified air temperature, featured by a moderately sized desiccant wheel, contributes to a more energy efficient DCS. Figure 6(b) - (c) document the energy performance of the MCCS and the MCMS with different membrane area and permeability. Increased total membrane plan area and increased water permeance both contribute to enhanced working-hour average system COP. As Equation (11) shows, both the total membrane plan area and water permeance are vary reciprocally with the pressure gradient for the same water vapor removal rate. Therefore, the impacts of the two parameters on the system performances are close to each other: increased values will reduce the pressure gradient across the membrane and thus increase the vacuum pressure on the permeate side. The rising vacuum pressure will reduce the vacuum pump work and contribute to a boosted efficiency. As the vacuum pressure approaches the water vapor partial pressure on the feed side, i.e. the pressure gradient approaches zero, further increases in either membrane plan area or water permeance have a diminishing return in enhancing the energy performance of membrane cooling systems, because the vacuum pump work is no longer the bottle-neck in the overall cooling work and the condensation work for the MCCS and the sensible cooling work for both the MCCS and the MCMS limit the energy performance. Results show that total membrane plan area of 1 m2/m2 (floor area) and water permeance of 6.2  10-7 mol/(m2 s) achieve the most economical and energy efficient membrane cooling systems for the studied case. In addition to the solid membrane dip-coated with polymers suggested by Bui et al. [26][27], it is reported a new form of the liquid membrane coated with hygroscopic liquid, e.g. TEG [31], PEG400 [32] and aqueous salt solutions (LiCl, LiBr, etc.) [33] can achieve higher water permeance of 6 × 10-6 mol/(m2 s Pa), an order of magnitude higher than that of solid membrane, with similar selectivity. It is reported that water vapor in the moist air can attach to the membrane surface much easier with hydrophilic liquid solutions on the feed side, and be rejected to the vacuum pump with a hydrophobic surface on the permeate side [31][33]. It will reduce the total membrane area for dehumidification, but it also requires periodic cleaning and replacement for maintenance since the mechanical strength of the liquid membrane is less robust than for solid materials. As Figure 6(d) shows, both membrane cooling systems achieve higher system COP with higher vacuum pump efficiencies. Inefficiencies of the vacuum pump will drastically increase the exit water vapor temperature (and enthalpy) given the pressure gap between the two ends of the pump, and therefore

17

increase both the vacuum pump work due to increased enthalpy differences, and the condensation work for the MCS to sensibly cool the vapor temperature to the condensation temperature. The increase in COP is not so significant at high efficiencies due to limitations of vacuum pressure and sensible cooling work. The result suggests using a vacuum pump with highest available isentropic efficiency to reduce the energy consumption of the membrane cooling systems. Figure 6(e) compares the energy performances of the MCCS with different stage pressure. As the stage pressure of the vacuum pump increases, the COP of the membrane system initially remains unchanged and then exponentially decreases with increasing pressure above 2 kPa. Although high stage pressure increases the condensation temperature and thus enhances the condensation efficiency with higher cooled water temperature, the non-isentropic compression process will substantially increase the water vapor temperature, adding a significant amount of both vacuum pump work and sensible cooling work on the system. The increase of sensible cooling load overwhelms the increase in sensible cooling efficiencies, and therefore results in lower system energy efficiencies. Since the pump work for the liquid water is negligible, the system performance does not benefit from the lower compressor work caused by the increased stage pressure. Figure 6(f) compares how the six DOAS cooling systems react with different heat recovery ratios. We can clearly see that energy performances of the DCS and the MCCS are most sensitive to the performance of the heat/enthalpy recovery wheels. The working-hour average COP of the DCS is boosted by 4.5 times as the sensible heat recovery ratio increases from 0.65 to 0.95. The MCCS also has almost 20% increase in the cooling energy efficiency. Since the HW transfers excess heat from the dehumidified air to the exhaust air, it reduces the sensible cooling work for the process air and the regeneration work for the exhaust air simultaneously. Thus, the energy efficiency of the DCS largely depends on the effectiveness of the heat recovery wheel. The ERW upstream of the MCCS preconditions the hot and humid outdoor air and reduces the sensible and latent cooling loads of the membrane system, which contributes to less work input on the vacuum pump and the GHE. The HCS and the ECS do not show much improvement in performance with increasing recovery ratios, because the chiller has to produce chilled water below the dew point temperature of the indoor air regardless of the preconditioning of the outdoor air, and the efficiencies of the chilled-based system are still limited by the compressor work between the evaporator and the condenser chambers.

5.4 Parametric Study In addition to the time-varying nature of their energy performance, different DOAS cooling systems also exhibit climate-specific performance. We performed parametric studies of the six systems to reveal their applicability and efficiency of cooling in different climates, and their sensitivity to indoor latent cooling part load ratio, defined as the actual indoor latent cooling load to the maximum designed value. Due to the high dimensionality in determining working conditions for the system, we limited our study to a fixed indoor air temperature of 24 °C and fixed indoor air relative humidity of 50%. For a fair comparison among different climates, equipment sizing and performances are maximized to remove their impacts on the system energy performances, and therefore the highest potentials of different systems are revealed. Figure 7 shows the energy efficiency of different DOAS cooling systems with a variety of annual average outdoor air temperature and relative humidity. We simulate cases with different indoor latent cooling part

18

load ratios to exhibit the energy efficiencies at different occupant activity intensities. The COP of the proposed DOAS systems increases with increasing outdoor temperature and relative humidity, due to the decreasing proportion of fan power in the overall work input. When the outdoor air condition is similar to the indoor air condition and the overall cooling load is very small, the fan power to ventilate the fresh air into the room through the duct system results in a very low energy efficiency. The ECS performs more energy-efficiently than the HCS, especially in hot and humid climates, due to its combined sensible and latent heat recovery. In most of the examined climates, especially hot and humid, the three nextgeneration DOAS cooling systems achieve COP of 15, more than twice of that for the BCS, which proves their uniformly superior energy performances in humidity removal relative to the industrial benchmark. For the MCMS, however, the large mass of water vapor in hot and humid climates will increase the pressure differences between the exit and the entrance of the vacuum pump to successfully transport the vapor to the bottom membrane, and thus the COP decreases in extreme hot and humid climates. As the indoor latent cooling load increases, the COP of the six proposed DOAS cooling systems does not change significantly, except for modest increases in cool and dry climates due to increased total cooling loads, and slight decreases in hot and humid climates due to increased total cooling work.

(a) Indoor latent cooling part load ratio = 0.1

(b) Indoor latent cooling part load ratio = 0.3

(c) Indoor latent cooling part load ratio = 0.9 Figure 7. COP of the proposed DOAS cooling systems at different working conditions when ground temperature is suitable for the use of GHE. Figure 8 compares the work input ratio of the proposed next-generation DOAS cooling systems and the baseline chiller system and its variations. The work input ratio is lower than one when the proposed system is more energy efficient than the BCS. As illustrated in the plots, the three next-generation cooling

19

systems always outperform the BCS and its variations, while reducing the total cooling work up to 80% at low indoor latent loads and 70% at high loads compared to the BCS, most notably in hot and humid climates. The MCMS performs similarly with the MCCS at low indoor latent loads, but is less energyefficient in hot and humid climates due to high vacuum pump work to reject large amounts of water vapor to the exhaust air stream. The HCS outperforms the BCS in hot and humid climates when large amounts of the sensible heat are removed by heat recovery, while the ECS consumes only half of the cooling work by the BCS in hot and humid climates due to enthalpy recovery.

(a) Indoor latent cooling part load ratio = 0.1

(b) Indoor latent cooling part load ratio = 0.3

(c) Indoor latent cooling part load ratio = 0.9 Figure 8. The cooling work ratio of the proposed next-generation DOAS cooling systems to the baseline chiller system at different working conditions when ground temperature is suitable for the use of GHE. As Figure 9 shows, we further compare the energy efficiencies of the next-generation DOAS cooling systems with the BCS when a chiller replaces the GHE. In tropical climates, the ground temperature is annually higher than the required chiller water temperature, and thus the GHE cannot provide sensible cooling for the membrane and desiccant systems. The results show that the substitute chiller reduces the energy efficiencies of the three next-generation systems around 15% compared with the GHE in most climates, especially for the DCS and the MCCS with large amounts of sensible cooling work (and condensation work for the MCCS). It also suggests climate-specific designs of energy-efficient DOAS

20

cooling systems to take full advantage of the energy reduction potential for each technology available in the region of interest.

(a) Indoor latent cooling part load ratio = 0.1

(b) Indoor latent cooling part load ratio = 0.3

(c) Indoor latent cooling part load ratio = 0.9 Figure 9. The cooling work ratio of the proposed next-generation DOAS cooling systems to the BCS at different working conditions when a chiller replaces the GHE.

6. Conclusion This paper reports the annual, climate-dependent performance of selected advanced outdoor-air cooling systems, obtained by combining detailed physical system models with annual building-load estimates, a significant advance over snap-shot, single-condition performance metrics of these systems. Thermodynamic models were developed to systematically compare the energy efficiencies of membrane cooling systems (MCCS and MCMS) and the desiccant cooling system (DCS), with the industrial benchmark, the baseline chiller system (BCS) and its variations (HCS and ECS), that widely used today. The comparison was made in real applications where the indoor occupant activities as well as the building envelope are included in the analysis. A low-energy office building prototype were constructed based on industry standards and flagship projects, which simulates the working conditions of next-generation DOAS cooling systems. A case study was conducted in Boston to reveal the time-dependent energy

21

performance for different cooling systems. A sensitivity analysis was performed to reveal the dependence of the system energy performances on individual equipment components. Climate-specific solutions were further suggested after running a parametric study of energy performances of different systems under varying climates. The key findings of this work include: 1. The proposed membrane cooling systems achieve higher energy efficiencies than the baseline chiller system and its variations in most examined climates. The advantages of the membrane cooling systems are separation of sensible and latent cooling by performing an isothermal water vapor separation and achieving high sensible cooling efficiency with renewable energy systems. Even in climates where ground temperature is not suitable for the GHE to work, the substitution of a chiller to provide the additional cooling still outperforms the baseline chiller system. The largest energy consumption in the membrane cooling system lies in collection of water vapor, including the vacuum pump work for both the MCCS and the MCMS and the condensation work for the MCCS. Increased membrane plan area and water permeance and a vacuum pump with high isentropic efficiency will contribute to enhanced system energy performances. 2. The proposed desiccant cooling system outperforms the baseline chiller system and its variations in energy efficiency in most examined climates, especially hot and humid conditions. Unlike the membrane cooling systems that are still under development for practical use, the components in the desiccant cooling system are available on the market today and can provide an intermediate energy-efficient system as the substitute of the baseline chiller system. A HW with high sensible recovery ratios significantly improves the system energy performance. Furthermore, solar thermal systems can also provide energy for regeneration work, making the system more competitive in comparison with the membrane cooling systems. 3. The ECS performs more efficiently than the HCS due to latent heat recovery, and both systems improve the energy efficiencies of the BCS with ―free‖ cooling from cool and dry exhaust air. The heat / enthalpy recovery wheel will be bypassed if the outdoor air is cooler / drier than the indoor air, which limits their applicability in mild or dry climates. 4. The GHE is energy efficient in providing additional sensible cooling downstream of latent cooling by membranes or desiccants but has limited applicability. The annual average ground temperature must be lower than the required chilled water temperature, which precludes its application in hot climates. Energy efficiencies are greatly reduced at high cooling loads when the chilled water flow rate is high, which suggests using more boreholes to reduce the sensible cooling work. 5. Climate-specific designs of next-generation DOAS cooling systems are necessary to achieve maximum energy efficiencies. The next-generation systems, the DCS, the MCCS and MCMS, are most energy efficient in warm humid and hot dry climates with moderate indoor latent cooling loads.

Acknowledgment This research is supported by the MIT Environmental Solutions Initiative. The authors also acknowledge helpful discussions with Prof. Leon Glicksman and Dr. Bui Thuan.

22

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[30] Lienhard, John H., and Leon R. Glicksman. Modeling and Approximation in Heat Transfer. Cambridge University Press, 2016. [31] Ito, Akira. "Dehumidification of air by a hygroscopic liquid membrane supported on surface of a hydrophobic microporous membrane." Journal of Membrane Science 175.1 (2000): 35-42. [32] Li, Jinlong, and Akira Ito. "Dehumidification and humidification of air by surface-soaked liquid membrane module with triethylene glycol." Journal of Membrane Science 325.2 (2008): 10071012. [33] Kudasheva, Alina, Yi‐ Jie Liu, and Akira Ito. "Aqueous salt solution liquid membranes, supported by nanoparticles, for water extraction from the atmosphere via air dehumidification." Journal of Chemical Technology & Biotechnology 93.10 (2018): 2851-2859. [34] Claridge, David E., and Charles H. Culp. "Systems and methods for multi-stage air dehumidification and cooling." U.S. Patent No. 8,641,806. 4 Feb. 2014. [35] Claridge, David E., and Charles H. Culp. "Systems and methods for air dehumidification and cooling with membrane water vapor rejection." U.S. Patent No. 8,500,848. 6 Aug. 2013. [36] Kneifel, J. "Prototype commercial buildings for energy and sustainability assessment: Design specification, life-cycle costing and carbon assessment." NIST Technical Note 1732 (2012). [37] Kneifel, Joshua D. "Prototype commercial buildings for energy and sustainability assessment: Whole building energy simulation design." Technical Note (NIST TN)-1716 (2011). [38] 2016 list of zero net energy buildings. New Building Institute. https://newbuildings.org/resource/2016-list-of-zne-buildings/ [39] 2018 Getting to zero status update and list of zero energy projects. New Building Institute. https://newbuildings.org/wp-content/uploads/2018/01/GTZ_2018_List.pdf [40] Crawley, Drury B., et al. "EnergyPlus: creating a new-generation building energy simulation program." Energy and buildings 33.4 (2001): 319-331. [41] Crawley, Drury B., et al. "EnergyPlus: energy simulation program." ASHRAE journal 42.4 (2000): 49. [42] Wetter, Michael, et al. "IEA EBC annex 60 modelica library-an international collaboration to develop a free open-source model library for buildings and community energy systems." Proceedings of building simulation 2015. 2015. [43] Wetter, Michael, et al. "Modelica buildings library." Journal of Building Performance Simulation 7.4 (2014): 253-270. [44] Florides, Georgios, and Soteris Kalogirou. "Ground heat exchangers—A review of systems, models and applications." Renewable energy 32.15 (2007): 2461-2478.

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[45] Sarbu, Ioan, and Calin Sebarchievici. "General review of ground-source heat pump systems for heating and cooling of buildings." Energy and buildings 70 (2014): 441-454. [46] Li, Min, and Alvin CK Lai. "Review of analytical models for heat transfer by vertical ground heat exchangers (GHEs): A perspective of time and space scales." Applied Energy 151 (2015): 178191. [47] Zeng, Heyi, Nairen Diao, and Zhaohong Fang. "Efficiency of vertical geothermal heat exchangers in the ground source heat pump system." Journal of Thermal Science 12.1 (2003): 77-81. [48] Yang, H., P. Cui, and Z. Fang. "Vertical-borehole ground-coupled heat pumps: A review of models and systems." Applied energy 87.1 (2010): 16-27. [49] Miyara, Akio. "Thermal performance and pressure drop of spiral-tube ground heat exchangers for ground-source heat pump." Applied Thermal Engineering 90 (2015): 630-637. [50] Eley, Charles. "Feasibility of ZNE by building type and climate." ASHRAE journal, July 2017: 32-37. [51] RESNET, January. "Mortgage Industry National Home Energy Rating System Standards." Residential Energy Services Network (2013). [52] ASHRAE, Standard. "Standard 62.1-2016 e Ventilation for Acceptable Indoor Air Quality, American Society for Heating." Refrigeration, and Air-Conditioning Engineers, Inc (2016). [53] Deru, Michael, et al. "US Department of Energy commercial reference building models of the national building stock." (2011): 1. [54] Jason Glazer PE, B. E. M. P. "Development of Maximum Technically Achievable Energy Targets for Commercial Buildings." ASHRAE Transactions 123 (2017): 32. [55] Griffith, Brent, et al. Assessment of the technical potential for achieving net zero-energy buildings in the commercial sector. No. NREL/TP-550-41957. National Renewable Energy Laboratory (NREL), Golden, CO., 2007. [56] Leach, Matthew, et al. Technical support document: Strategies for 50% energy savings in large office buildings. No. NREL/TP-550-49213. National Renewable Energy Laboratory (NREL), Golden, CO., 2010.

26

Appendices A. Electric Chiller Performance Model [10] The model uses polynomial interpolation on the performance curves of a variety of water-cooled electric chillers on the market. Given the maximum cooling capacity of the chiller and the reference working condition with a nominal COP value, the actual performance is adjusted by part-load ratio, a function of leaving chilled water temperature and entering condenser temperature. The two temperatures reflect the working temperatures of the evaporator and condenser, determining the refrigerant pressure difference between the two heat exchangers and thus the total compressor work in the system. Therefore, we can express the operational work input for the electric chiller as follows: (16) where the subscript cap refers to the maximum capacity; ref indicates the reference condition; c denotes cooling; EIRfT, EIRfPLR, CAPfT are three multipliers indicating adjustment from reference conditions; EIR denotes energy input to cooling output ratio, the inverse of the COP value; and PLR is part load ratio, defined as the actual cooling load divided by the chiller’s available cooling capacity. ) ̇ ( (17) EIRfT and CAPfT are both expressed by biquadratic performance curves parametrized with leaving chilled water temperature (Unit: °C) and entering condenser temperature (Unit: °C), while EIRfPLR is expressed by a quadratic performance curve parametrized with PLR [10]. (18) (19) (20) where C, D and E’s are all coefficients of the performance curve, specified by experimental tests for different brands of the electric chiller. Table A lists the best performer of available water-cooled electric chillers with data stored in EnergyPlus based on the highest reference COP. In order to maximize the efficiency of the chiller, we optimally split the oversized chiller according to the end-use peak loads. Table A. Performance curve coefficients for electric chiller (Trane CVHF 2567kW/11.77COP/VSD) Parameter

Value 4.561990×10-1 -3.333063×10-2 -1.763576×10-3 5.624520×10-2 -8.535085×10-5 3.567604×10-4

Parameter

Value 4.549428×10-1 2.748161×10-2 -4.298342×10-3 4.363928×10-2 -1.352509×10-3 2.964225×10-3

Parameter

B. Parametrization of Desiccant Wheel Table B. Non-dimensional parameterization of the desiccant wheel model [18]

27

Value 1.039419×10-1 4.184975×10-1 4.774210×10-1

Category

Parameter

Expression

Physical explanation Temperature Humidity ratio

Variables

Time Transfer capacity Physical parameters of desiccant wheel

(

)

(

)

(

(

(

) ̅̅̅ ̅̅̅ (

̅̅̅

Sorption capacity

)

Gradient of relative humidity line in the psychrometric chart Gradient of water content in the desiccant with respect to humidity ratio Average Gradient of water content in the desiccant with respect to temperature

)

(

Thermal capacity

)

) ⁄

Coefficients (

⁄ )

(

)

( (

)( )

(

))

(

)(

(

))

C. Thermodynamic Modeling of Ground Heat Exchanger A ground heat exchanger, also known as geothermal heat exchanger (GHE), exploits seasonal geothermal energy storage to heat or cool the process air. Experimental studies have shown a relatively constant ground temperature throughout a year below a certain depth due to the thermal inertia of soil [44]. The GHE is used in the system to exchange heat with the ground, and in other cases couple with a heat pump to form a residential and commercial HVAC system [45]. The design of GHE could be open-loop or closed-loop, horizontal or vertical oriented [44][46]. In our analysis below, we consider a closed-loop vertical borehole system for sensible cooling. The closed-loop GHE typically includes a heat exchanger above ground, with the working fluid connecting the heat exchanger with the borehole system buried in the ground. The borehole consists of

28

small-diameter U-shaped pipes with grouting materials to fill the borehole, as shown in Figure 10. Water is usually used as the working fluid that runs through the pipes and exchanges heat with surroundings.

(a) (b) Figure 10. Schematic of GHE. (a) refers to the closed-loop vertical borehole GHE; (b) is a horizontal section A-A of the underground borehole. The grey area in the borehole denotes the grout. The red/blue area in the pipe refer to the working fluid with hot/cold temperature. H denotes the depth of the pipe; D denotes the distance between the pipe centers; dp denotes the diameter of the pipe; db denotes the diameter of the borehole. A constant effectiveness model is used to evaluate the inlet working fluid temperature effectiveness of the heat exchanger, ε, is taken as 0.75. ( ) ̇ ( ̇ ( ̇

)

) { ̇

( ̇

. The

)

(21)

}

(22)

where and refer to the inlet and outlet air stream temperature, respectively and chilled working fluid temperature.

is the inlet

We can treat the borehole system as a storage-type heat exchanger [47] and similarly define another effectiveness, , to evaluate the potential of flushing the heat wave to the ground. (23) Here the thermal capacity of the ground is much larger than the working fluid, and thus the effectiveness model could be simplified as the expressions above. Note that is the borehole wall temperature. To determine the effectiveness of the ground heat exchanger, we make some assumptions to reasonably simplify the analysis. 1) The ground is treated as a constant temperature sink with infinite thermal capacity; 2) dimensional analysis has shown that the heat flow is approximately steady-state with hourly simulation frequency [46], and conduction is dominant in the borehole cross-section because the borehole depth is far greater than its diameter; 3) heat flow in the grout and pipe walls is assumed to be negligible in the radial direction. Zeng et al. [47] developed a quasi-three-dimensional model by equating the axial temperature gradient along the borehole depth to radial conduction inside the pipe and thermal interferences between injection and ejection pipes.

29

̇ (24) {

̇

where index 1 and 2 refer to temperature profile in the injection and ejection pipe, respectively. The boundary conditions of the ODEs are, {

(25)

The temperature profile is solved by Laplace transformation and the thermal resistance in the borehole cross-section is derived by a Kelvin line source model [48][49]. Thus, the analytical expression of the effectiveness of the ground heat exchanger can be obtained as, ( ) (26) ( ) √

̇

√( * (

)(

(27)

)

)

(

)+ (28)

* (

)

(

)+

where the borehole thermal conductivity depends on the pipe wall and grouting materials, an average estimated conductivity is given as 2.0 W/m K, and the ground thermal conductivity is approximated as the thermal conductivity of clay, i.e. = 1.2 W/m K. When we plug in typical geometrical data for a borehole system ( = 0.025 m, = 0.1 m, = 0.05 m, = 50 m), = 0.70. Notice that is not sensitive to a large value ( > 3 is satisfied with ̇ < 0.2 kg/s) due to the saturation nature of the hyperbolic tangent function. Furthermore, when a borehole system is designed, is also fixed. So it is without loss of generality to assume that the effectiveness of the ground heat exchanger is constant with different working conditions, and herein we take the value as 0.70. The borehole wall temperature is determined by equating the radial heat flux from inside to outside of borehole. ( ⁄ ) (29) ⁄ (30) * ( where and respectively,

)

+

(31)

represent the thermal resistance per length of the pipe for the borehole and ground, is the thermal diffusivity of the ground, and is the working hours of the GHE. Note that

30

the expression of is derived by an infinite cylindrical-surface source model [48][49]. For a given climate, the annual average ground temperature is fixed. As discussed above, the GHE design parameters determine the thermal resistance of the borehole and ground, and it does not change with different working conditions. Therefore, the borehole wall temperature can be calculated by the average working fluid temperature in the pipe. After deriving an energy equation for the air-water heat exchanger above ground, the system is then closed. The inlet and outlet chilled water temperature as well as its mass flow rate will then be determined. The work done by the GHE is through the pump work that drives the water running through the pipe with positive pressure gradient. The pump work is then calculated according to the Darcy-Weisbach equation with air properties changed for those of the working fluid [29][30].

D. Low-Energy Building Design Specifics Table D. Low-energy office building prototype Category General Loads

HVAC

Domestic Hot Water

Insulation

Item Occupant density Lighting power density Lighting illuminance level Equipment power density Mechanical ventilation rate Sensible / Latent heat recovery rate Infiltration rate Supply temperature Peak flow rate Windows Ground floor Slab Partition Façade (wall) Roof

Occupant density Schedules Lighting

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Value (Unit) 0.1 (p/m2) 2.6 (W/m2) 500 (lux) 8.0 (W/m2) 0.55 (L/s/m2) 0.75 0.6 (/hr) 60 (°C) 0.1 (L/h/m2) U-1.3 (W/m2K) R-3.5 (m2K/W) R-0.9 (m2K/W) R-0.9 (m2K/W) R-4.4 (m2K/W) R-6.2 (m2K/W)

Equipment

HVAC

Domestic hot water

Notes: 1. The occupant density for office buildings is referenced to the 16 representative commercial office buildings in 16 cities representative of U.S. climate zones, according to the survey from DOE research [53]. According to the statistics in US Census Bureau, we chose the median size of commercial building, 1760 m2, as the representative floor area of the low-energy office building prototype. 2. The lighting and equipment power densities are both referenced to an ASHRAE report [54], stating the maximum technically achievable energy targets for commercial building. The ―Max Tech‖ quoted in this paper represents the highest efficiency of devices that either currently available or can be reasonably expected to be available by 2030 by at least two manufacturers. The lighting power density target can be achieved by using high-efficient LED on both exterior and interior lighting system. According to the state-of-the-art efficacy of the LED lighting, the lighting power density is suggested by a reduction factor of 0.3208 on the up-to-date ASHRAE Standard 90.1-2016. By comparison, Griffith [55] also created a building energy model in an earlier research, based on thousands of existing buildings from the 2003 Commercial Building Energy Consumption Survey (CBECS), categorized with different building types. The lighting power density given in the paper is 5.5 W/m2, 50% of the value suggested by ASHRAE Standard 90.1-2004 [56]. The 50% reduction in lighting energy consumption can be achieved by integrating daylighting controls on the lighting systems. However, it is still almost twice of the Max Tech target, which shows a great potential to reduce the lighting power consumption in building operations.

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3.

4. 5. 6. 7.

8.

The equipment power density target is determined by an additional 25% decrease on plug and miscellaneous loads of existing buildings in the CBECS dataset. The decrease in equipment power usage comes from power management on machines and desktops, automated idle mode at rest, and replacement of desktops, laptops and printers with high-efficiency models [54][55]. Similar to the settings of the low-energy residential building prototype, the mechanical ventilation rate is determined by the minimum fresh air intake rate required per area and per person, default as 0.3 L/s/m2 and 2.5 L/s/people, respectively, regulated in ASHRAE Standard 62.1-2016 Table 6.2.2.1 [52]. The heat recovery rate we chose conforms to the same model used in low-energy residential buildings. The infiltration rate for the office buildings is referenced to the ASHRAE Standard 90.1-2016 [9]. The domestic hot water supply temperature and peak flow rate conform to the same settings in low-energy residential buildings. The insulation construction of commercial buildings depends on specific climates, with R values stipulated in ASHRAE Standard 90.1-2016 [9]. The windows of the building are assumed to have triple-layer low-emissivity 3-mm clear glass (thermal conductivity: 0.9 W/m K) sandwiched with 6 mm Argon (thermal conductivity: 0.016 W/m K). The window-to-wall ratio is assumed 0.4 in the north and west, and 0.6 in the south and east. The ground floor of the building includes four layers of materials, quantified by thickness: 22.5 cm gypsum fiber board (thermal conductivity: 0.32 W/m K), 7.5 cm rigid foam (thermal conductivity: 0.03 W/m K), 10 cm reinforced concrete (thermal conductivity: 2 W/m K), and 20 cm rammed earth (thermal conductivity: 0.75 W/m K). The concrete slabs for each floor of the building include three layers of materials: double-layer 1.5 cm impact sound insulation (thermal conductivity: 0.035 W/m K), and one inner layer of 10 cm general concrete. Similar to the horizontal slabs, the vertical partitions of the building include three layers of materials: double-layer 1.5 cm impact sound insulation, and one inner layer of 5 cm general concrete. The façade walls of the building include five layers of materials: 1.3 cm oriented strand board (thermal conductivity: 0.13 W/m K), 9.5 cm XPS board (thermal conductivity: 0.034 W/m K), 10 cm reinforced concrete, 8 cm foam glass, and 1.3 cm gypsum fiber board. The roof of the building includes four layers of materials: 2 cm asphalt high binder content (thermal conductivity: 0.75 W/m K), 15 cm rigid foam with aluminum coating (thermal conductivity: 0.025 W/m K), 10 cm reinforced concrete, and 2 cm oriented strand board. The schedules of building operations follow the instructions in commercial office building models researched by Department of Energy (DOE) [53].

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Conflict of Interest Statement

We have no conflict of interest to declare. This statement is to certify that all Authors have seen and approved the manuscript being submitted. We warrant that the article is the Authors' original work. We warrant that the article has not received prior publication and is not under consideration for publication elsewhere. On behalf of all Co-Authors, the corresponding Author shall bear full responsibility for the submission.

This research has not been submitted for publication nor has it been published in whole or in part elsewhere. We attest to the fact that all Authors listed on the title page have contributed significantly to the work, have read the manuscript, attest to the validity and legitimacy of the data and its interpretation, and agree to its submission to Energy and Buildings.

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