Net-zero Nation: HVAC and PV Systems for Residential Net-Zero Energy Buildings across the United States

Abstract

This study compared the energy performance and initial cost of photovoltaic (PV) and heating, ventilating, and air-conditioning (HVAC) equipment for a residential net-zero energy building (NZEB) in different climate zones across the United States. We used an experimentally validated building simulation model to evaluate various electrically-powered and commercially-available HVAC technologies. The HVAC accounted for 23.8 % to 72.9 % of the total building energy depending on the HVAC option and climate zone. Each HVAC configuration was paired with a PV system sized to exactly reach the net-zero energy target, so the economics were compared based on the initial PV + HVAC cost. Mechanical ventilation was considered with and without heat recovery; the heat recovery ventilator (HRV) saved a significant amount of energy in cold winter months and hot summer months, and the energy recovery ventilator (ERV) provided additional benefit for humid zones. The HRV was cost-effective in the cold northern latitudes of Chicago, Minneapolis, Helena, and Duluth, where energy savings reached 17.3 % to 19.7 %. In other climates, ventilation without recovery was more cost effective, by 1 % to 9 %, and sometimes even more energy efficient. The ERV was never the lowest cost option. A ground-source heat pump (GSHP) and an air-source heat pump (ASHP) were compared, with the GSHP providing significant energy savings, 24.3 % to 39.2 %, in heating-dominated climates (Chicago through Duluth). In warmer climates, the GSHP saved little energy or used more energy than the ASHP. The PV + HVAC cost was lower everywhere with the ASHP, though it is possible for colder climates that a carefully sized GSHP and ground loop could be cost-competitive. The energy and cost data as well as the required PV capacity could guide HVAC and PV designs for residential NZEBs in different climate zones.

1. Introduction

Residential buildings accounted for 21 % of the total building energy consumption and 20 % of the total carbon dioxide emissions in the U.S. in 2016 [1]. Therefore, improving energy efficiency in residences is an important national priority and could significantly reduce greenhouse gas emissions [2]. Net-zero energy buildings (NZEBs) are defined for this study as producing at least as much (site) energy as they consume over a given year [3–5]. The U.S. Department of Energy aims to achieve “marketable zero energy homes in 2020 and commercial zero energy buildings in 2025” [6]. The American Society of Heating, Refrigerating, and Air-Conditioning Engineers (ASHRAE) set a goal of market-viable NZEBs by 2030 [7].

Renewable energy resources for NZEBs have been extensively studied and compared [8–11], while heating, ventilating, air-conditioning (HVAC) systems have been investigated less. Even with advanced envelope configurations featuring very airtight construction and enhanced insulation, HVAC still comprises the largest share of energy consumption in these buildings [12, 13], more than 40 % [14]. Applying advanced HVAC technologies to NZEBs is therefore an important strategy to further improve their energy performance.

There have been a number of studies on alternative HVAC technologies for residential NZEBs. Ng and Payne [15] evaluated the energy consumption of an air-source heat pump (ASHP) with and without a sensible heat recovery ventilator (HRV) for a residential NZEB in Maryland, U.S. The impact of the HRV on total HVAC energy ranged from 36 % savings in winter to a 5 % increase in mild seasons. Marszal and Heiselberg [16] compared the life-cycle cost of a residential NZEB in Denmark with HVAC options including a solar heat pump, a ground-source heat pump (GSHP), and a district heating system. They found the solar heat pump was the most energy efficient option. Mohamed et al. [17] investigated the performance of a residential NZEB in Finland based on four metrics, including primary energy, site energy, equivalent CO₂ emission, and energy cost. They considered different energy systems, including five heating options, i.e., electric heating, district heating, GSHP, light oil boiler and wood pellet boiler, showing that GSHPs could greatly reduce the PV areas required to achieve NZEB. Norton and Christensen [18] described a case study of an affordable residential NZEB in a cold climate (Denver). Among the four heating options (solar heating, GSHP, natural gas furnace, and electric resistance heating), the natural gas furnace was the most cost effective.

The local weather characteristics, including solar radiation, ambient temperature and relative humidity, have great influence on PV generation, building thermal load, and HVAC performance; therefore, the most appropriate NZEB designs vary by climate zone. Deng et al. [19] discussed building performance for typical construction in Shanghai (humid) and Madrid (dry), with advanced technologies including solar-assisted ASHP, heat recovery, evaporative cooling, and night radiant cooling. The building in Shanghai consumed 9.8 % more electricity but produced 25.4 % less electricity compared to a building in Madrid. Elkinton et al. [20] investigated a residential NZEB in five U.S. climate zones, but mainly focused on how the designs and economics were affected by renewable resources, energy prices, and state incentives (e.g., net metering). They found that for locations with large heating loads and low utility prices, the NZEB was more expensive. Kapsalaki et al. [21] presented a method to achieve economically efficient residential NZEB designs in three European climates: Stockholm (cold winter), Lisbon (mild winter), and Iraklion (mild winter but warm summer). However, the study focused on the economically-efficient design of NZEBs at the early design stage, without analysis of the detailed operation performance of various HVAC systems. Kneifel and O’Rear [22] evaluated both the energy and economic performance of a NZEB for 45 U.S. locations in a mixed-humid climate zone. They concluded that the NZEB economy improved the further south in that climate zone it was located.

This literature review indicates that the existing studies on HVAC systems of NZEBs covered limited options and climates. This paper studied various commercially-available (i.e. representative of current consumer choice) HVAC options for residential NZEBs applied in all the climate zones across the contiguous U.S. We carried out minutely transient building energy simulations, using an experimentally validated model, for ventilation and heat pump options in 15 representative cities. The HVAC options were each paired with a PV system with capacity to exactly meet the net-zero energy target. This approach highlighted the cost trade-off between NZEB designs with higher efficiency HVAC equipment vs. larger PV arrays. The options were compared in terms of energy performance and total initial cost of the HVAC + PV system. The cost-effectiveness of the options was compared solely on the initial cost since there was no energy cost (operation cost) for all the options. The objective of this study was to identify the relative merits of alternative HVAC technologies and provide recommendations for HVAC and PV designs for residential NZEBs in different climate zones.

2. Description of the residential NZEB

Fig. 1 shows the single-family residential NZEB (Gaithersburg, MD, U.S.) that was the basis of the experimentally verified model presented here. The house consisted of two stories of living area (251 m²), an attic (within the envelope), and a conditioned basement (135 m²). The first floor included the kitchen and dining room, a family room, an office, and a full bathroom. The second floor included a master bedroom with adjoining bathroom, two additional bedrooms, and two bathrooms.

Fig. 1.

Residential NZEB on NIST campus

The house was unoccupied, but the activities associated with a family of four were emulated by computer-activated appliances, plug loads, lighting, water draws, and devices per weekly schedules specified in [23]. These schedules were based on standard user profiles developed for the Building America program [24]. During the first year of operation, the residential NZEB generated approximately 484 kWh more electricity than it consumed, exceeding the net-zero energy goal [25].

The NZEB featured a wide array of energy efficient technologies and equipment. To reduce the heating and cooling loads, the thermal envelope employed increased insulation, double-paned windows, and very airtight construction (0.63 air change per hour at 50 Pa measured by a blower-door test). Table 1 presents the main envelope specifications, with more detailed information and performance data provided in [25–27].

Table 1.

Main envelope specifications of the residential NZEB

Envelope	Brief specification
External wall	• 5.1 cm × 15.2 cm wooden studs on 61 cm centers
	• 1.3 cm plywood sheathing; air and moisture barrier
	• Two layers of 5.1 cm foil-faced polyisocyanurate foam board
	• 2.5 cm × 10.1 cm wood furring strips secured to the framing
	• Blow-in cellulose insulation in cavities
	• U-factor: 0.13 W/(m2·°C)
Windows	• Double-hung insulated fiberglass frames
	• Two panes of low-e coated glass with a suspended film
	• Cavities filled with an inert gas
	• U-factor: 1.14 W/(m2.°C)
	• Solar heat gain coefficient: 0.25
Roof	• 1.3 cm and 1.6 cm plywood sheathing; air-moisture barriers
	• 12.7 cm foil-faced polyisocyanurate insulation; asphalt shingles
	• Blown-in cellulose insulation in rafter cavities
	• Interior finished using 1.3 cm thick gypsum board
	• U-factor: 0.08 W/(m2·°C)
Basement wall	• 25.4 cm poured concrete wall with 10.2 cm stem wall
	• Damp proofing; 5.1 cm extruded polystyrene board
	• 5.1 cm foil-faced polyisocyanurate insulation with 1.3 cm gypsum board
	• U-factor: 0.24 W/(m2·°C)
Basement floor	• Open cell spray foam insulation in floor joist cavities
	• 10.2 cm concrete slab; 0.15 mm polyethylene vapor barrier
	• 5.1 cm extruded polystyrene insulation
	• U-factor: 0.56 W/(m2·°C)
Window-to-wall ratio	• 0.13
Airtightness	• 0.63 air change per hour at 50 Pa, measured by a blower-door test

The house generated electricity using thirty-two 320 W positively-grounded monocrystalline silicon photovoltaic (PV) modules, with nominal 19.6 % efficiency [25], providing a maximum power of 10.24 kW DC at standard test conditions (normal irradiance at 1000 W/m², module cell temperature at 25 °C) [28]. Two inverters with a nominal efficiency of 95.5 % were used to maximum-power-track the modules and convert their direct current to 60 Hz alternating current. The excess electricity was exported to the utility grid.

The HVAC system consisted of an ASHP and a heat recovery ventilator (HRV) (Fig. 2). The ASHP was a split-system with three heating stages, two cooling stages, and a dedicated dehumidification mode, with the performance parameters provided in model section (section 3.1, Table 3).

Fig. 2.

Schematic diagram of the HVAC system

Table 3.

Components and parameters of different subsystems of the baseline TRNSYS model

Subsystem	Component	TRNSYS type	Main parameters
Building	Building	Type 56	living area: 251 m2; basement area:135 m2
	Air infiltration	Type 932	effective leakage area: 244 cm2 (@ 50 Pa)
PV	PV array	Type 194b	module area: 32×1.472 m2; orientation: south tilt angle: 18.4°; maximum power: 10.24 kW DC inverter efficiency: 95.5 %; frequency: 60 Hz
HVAC	1st and 2nd stage ASHP	Type 922	1st stage rated airflow rate: 1060 m3/h 1st stage rated capacity: 5425 W total cooling, 4088 W sensible cooling, 5180 W heating 1st stage rated power: 1440 W cooling, 1398 W heating
			2nd stage rated airflow rate: 1420 m3/h 2nd stage rated capacity: 7067 W total cooling, 5360 W sensible cooling, 8041 W heating 2nd stage rated power: 2420 W cooling, 2126 W heating
	3rd stage ASHP	Type 954	3rd stage rated airflow rate: 1420 m3/h 3rd stage rated heating capacity: 8041 W 3rd stage rated heating power: 2126 W supplemental electric resistance heating: 10 000 W
			1st stage rated airflow rate: 1025 m3/h 1st stage rated cooling capacity: 5997 W total, 4177 W sensible
	Dehumidification ASHP	Type 922	1st stage rated cooling power: 1820 W 2nd stage rated airflow rate: 933 m3/h 2nd stage rated cooling capacity: 1653 W total, 423 W sensible 2nd stage rated cooling power: 1230 W
	HRV	Type 667b	outdoor air: 195 m3/h; defrost temperature trigger: −10 °C sensible effectiveness: 0.72; latent effectiveness: 0.01
	HRV fans	Type 111b	(each of the two fans) rated flow rate: 195 m3/h rated power: 27 W; motor efficiency: 0.90
DHW	Solar collector	Type1b	collector area: 2.1 m2; intercept efficiency: 0.744 1st order efficiency coefficient: 3.6707 W/(m2-K) 2st order efficiency coefficient: 0.00543 W/(m2·K2)
	Solar storage tank	Type 534	volume: 0.189 m3; height: 1.143 m loss coefficient: 0.59 W/(m2·K)
	Heat exchanger	Type 91	heat exchanger effectiveness: 0.44
	Pumps	Type 114	rated flow rate: 196 kg/h glycol, 999 kg/h water rated power: 80 W glycol, 80 W water
	HPWH	Type 938	total airflow rate: 765 m3/h; blower power: 5 W rated heating capacity: 2025 W rated heating power: 794 W
	Electric heater	Type 1226	heating capacity: 3600 W; thermal efficiency: 1.0
	HPWH tank	Type 534	volume: 0.2953 m3; height: 1.5939 m loss coefficient: 1.0 W/(m2·K)
	Pipes	Type 31	diameters, lengths, and loss coefficients in [31]

The heat pump operated according to the thermostat temperature- and time-triggers listed in Table 2. In cooling and heating, the heat pump started in 1^st stage, then advanced to 2^nd stage when either the temperature- or time-trigger criteria were met. A 3^rd stage of heating activated after 40 min of 2^nd stage operation, or if the temperature was 3.3 °C from the setpoint; the compressor operated the same as on 2^nd stage (i.e. at the same capacity and COP) and was supplemented by 10 000 W of electric resistance heat. The heat pump dehumidification modes operated when the relative humidity was over the setpoint but the temperature setpoint was satisfied. For 1^st-stage dehumidification, the heat pump operated in the cooling mode with the compressor in 1^st stage and a reduced fan speed. In 2^nd-stage dehumidification (i.e., dedicated dehumidification mode) the compressor operated in 1^st stage and the supply air was reheated (to avoid overcooling) using a portion of hot refrigerant vapor that bypassed the condenser. The defrost cycle of ASHP was activated after 90 min of accumulated compressor runtime while the outdoor temperature was below 1.7 °C.

Table 2.

Heat pump control logic of the residential NZEB

Mode	Setpoint	Turn-on trigger	1st stage	2nd stage	3rd stage
Heating (°C)	20.5	Deadband (°C)	0.1	1.1	3.3
		Time delay (min)	-	10	40
Cooling (°C)	23.9	Deadband (°C)	0.2	2.8	-
		Time delay (min)	-	40	-
Dehumidification (%)	48	Deadband (%)	2	-	-
		Time delay (min)	-	6	-

The HRV recovered sensible heat from the exhaust air to pre-condition the supply outdoor air, at an airflow rate of approximately 195 m³/h (0.15 air change per hour) based on ASHRAE Standard 62.2 [29] and the manufacturer’s specifications. The HRV switched to a defrost mode when the outdoor air temperature fell below −10 °C. During defrost, the unit recirculated the indoor air for 7 min, then returned to normal operation for 22 min.

Fig. 3 shows the domestic hot water (DHW) system with an indirect solar hot water (SHW) subsystem and a heat pump water heater (HPWH) subsystem. The water was first preheated by the SHW subsystem and then heated to the setpoint by the HPWH, if necessary, before being supplied to the end use. Water from the SHW tank was tempered to a maximum of 49 °C before entering the HPWH tank. After exiting the HPWH tank, the hot water was delivered to the fixtures and mixed with cold water to achieve use-specific temperatures including: 41 °C for sinks and showers, 43 °C for baths, and 49 °C for the dishwasher and clothes washer.

Fig. 3.

Schematic of the DHW system

3. Methods

We studied various HVAC technologies in different climate zones using a validated model of the residential NZEB. This section describes the NZEB model, including the baseline HVAC and PV configurations for the test house in Gaithersburg. Next, we present the representative cities in different climates, HVAC sizing and options for each climate, and the evaluation indices.

3.1. Model of the residential NZEB and baseline HVAC and PV systems

A TRNSYS [30] model of the building and baseline mechanical equipment was created to simulate the test house in Gaithersburg (Fig 1). The model is described briefly here; more detail, including the validation results, can be found in the appendix and references [31, 32]. Table 3 presents the model parameter configurations of the PV, HVAC, and DHW subsystems. In the alternative climates discussed in section 3.2, the heat pump and PV sizes were adjusted to meet the building load and electricity generation requirements, and the model parameters were adjusted accordingly. The DHW system is the same for all climates.

We used the measured capacity and power data to formulate a TRNSYS performance map for each mode of the baseline ASHP [14]. The performance map corrects for outdoor temperature, indoor dry-bulb and wet-bulb temperatures, and indoor airflow rate. Fig. 4. shows the performance map capacity and power as functions of outdoor temperature for the noted indoor conditions.

Fig. 4.

Capacity and power of the 7.1 kW (2 ton) ASHP vs. outdoor air temperature

We linearly correlated the ASHP’s measured electrical energy (kWh) per defrost cycle to the ambient air temperature (°C) according to Equation (1):

$$E_{defrost}=-0.0186t_{air}+0.4005$$ (1)

The solar thermal collector efficiency in TRNSYS is presented in Equation (2):

$${\eta }_{collector}=a_0-a_1\frac{t_{fluid}-t_{air}}{G}-a_2\frac{{\left(t_{fluid}-t_{air}\right)}^2}{G}$$ (2)

where t_fluid is the average heat transfer fluid temperature in the collector (°C); t_air is the ambient air temperature (°C); G is the total solar radiation incident on the surface (kJ/(h·m²)); a₀, a₁, and a₂ were respectively the intercept efficiency, the 1^st order efficiency coefficient (W/(m²·K)), and the 2^st order efficiency coefficient (W/(m²·K²)), from the manufacturer, Table 3.

We used a time step of 1 min to capture effects related to the equipment controls, which operated at various time scales. We previously validated the TRNSYS model with a year’s worth of measured data [26]. The measurement uncertainty associated with the HVAC system, as well as all other electrical/mechanical subsystems within the NZEB were described in detail [27]. The simulated and measured building thermal loads agreed within 4 % for heating and within 5 % for cooling (Fig 5(a)). The subsystem energy consumptions agreed between 0.9 % and 7.7 % (Fig. 5(b)). The PV generation matched within 3.1 % for the annual total, and within an average of 6.6 % for the monthly totals (Fig 5(c)). The high deviation in February and March (Fig 5(c)) was caused by snow covering the PV panels during operation of the test house, which was not included in the model. The monthly ASHP energy consumption matched within 13 % for cooling and 10 % for heating (Fig. 5(d)). Finally, the inlet and outlet water temperatures of the SHW tank and HPWH tank also matched closely (Fig. 5(e) and (f)). The remainder of section 3 describes how the validated model was used to investigate the studied HVAC options in different climate zones.

Fig. 5.

Validation of TRNSYS models for the residential NZEB

3.2. Representative cities in different climate zones

Per ASHRAE Standard 90.1 [33], U.S. climate zones are classified according to the heating and cooling degree days and ambient moisture levels. The U.S. includes eight major climate zones, where zone 1 has the most cooling degree days and zone 8 has the most heating degree days. Some of the major climate zones are further divided into humid, dry, and marine climate types (A, B, and C), leading to 15 total climate zones. We evaluated the NZEB and the HVAC options in one city from each of the 15 climate zones (Fig. 6 and Table 4). Note that we didn’t consider zone 8 (Alaska), and considered two cities (Los Angeles, CA and Las Vegas, NV) in zone 3B. The hourly weather files used in TRNSYS were the TMY2 (Typical Meteorological Year) data sets derived from the 1961–1990 National Solar Radiation Data Base [34].

Fig. 6.

Representative cities in different U.S. climate zones

Table 4.

Weather features of the representative cities

Climate zone	Representative city	Solar radiation (MJ/m2)	Temperature (°C) *			Relative humidity (%) *
		Cumulative	Max	Min	Avg	Max	Min	Avg
1A	Miami, FL	6453	33.4	3.9	24.3	100.0	22.5	72.5
2A	Houston, TX	5826	36.1	−9.7	20.0	100.0	16.0	75.3
2B	Phoenix, AZ	7620	46.1	−2.8	22.5	99.5	4.0	36.3
3A	Atlanta, GA	6124	36.1	−11.1	15.9	100.0	13.5	68.3
3B	Los Angeles, CA	6545	33.9	4.4	16.7	100.0	6.0	70.0
3B	Las Vegas, NV	7478	44.2	−4.2	19.5	94.0	3.0	29.5
3C	San Francisco, CA	6256	34.7	0.3	13.1	100.0	19.5	73.7
4A	Baltimore, MD	5335	35.6	−15.6	12.6	100.0	19.0	67.1
4B	Albuquerque, NM	7369	37.0	−10.9	13.2	100.0	2.5	45.3
4C	Seattle, WA	4398	36.7	−4.7	10.9	100.0	17.0	73.2
5A	Chicago, IL	5153	35.0	−22.5	9.8	100.0	17.5	69.1
5B	Boulder, CO	6078	36.4	−23.1	9.9	100.0	4.0	51.9
6A	Minneapolis, MN	5209	34.7	−29.2	7.3	100.0	18.5	69.4
6B	Helena, MT	5175	35.9	−29.2	6.8	100.0	12.0	56.6
7	Duluth, MN	4862	30.9	−31.7	3.2	100.0	19.5	71.3

^*on an hourly basis

3.3. HVAC options

We evaluated different HVAC options for the ventilation and heat pump subsystems. Although there are many HVAC options, this study focused on commercially- and widely-applicable options. For each subsystem study, the equipment in the other subsystems was held constant: (1) for ventilation options, we compared three options with the same heat pump subsystem; (2) for heat pump options, we compared two options with the same ventilation subsystem.

(1). Ventilation options

We compared a mechanical ventilation system (a) without heat recovery, (b) with an HRV, and (c) with an ERV (energy recovery ventilator, with sensible and latent recovery). We fit the manufacturer’s specifications for the sensible effectiveness (ε) and the power (P) as a function of the airflow rate (m):

$$\epsilon ={\epsilon }_0\left[c_0+c_1\frac{m}{m_0}+c_2{\left(\frac{m}{m_0}\right)}^2\right]$$ (3)

$$P=P_0\left[c_0+c_1\frac{m}{m_0}+c_2{\left(\frac{m}{m_0}\right)}^2\right]$$ (4)

where ε₀, P₀ and m₀ are the rated sensible effectiveness, power, and mass flow rate; c₀, c₁, and c₂ are the coefficients (Table 5). Note that ventilation without recovery only used a supply fan (27 W), whereas the HRV (54 W) and ERV (60 W) each use two fans. The ventilation system affects the heat pump energy use due to its impact on heating and cooling loads, so the options were compared based on the combined heat pump and ventilation energy. For each ventilation option, the ASHP was used in the simulation.

Table 5.

Correlation parameters of different ventilation options

Parameter	No heat recovery	HRV	ERV
Rated sensible effectiveness	0.00	ε0 = 0.72	ε0 = 0.68
Sensible effectiveness coefficients	Not Applicable	c0 = 1.273	c0 = 1.103
		c1= −0.3246	c1= 0.0162
		c2= 0.0521	c2= −0.119
Latent effectiveness	0.00	0.01	0.47
Airflow rate (m3/h)	195	195	195
Rated power (W)	P0 = 27.0	P0 = 54.0	P0 = 60.0
Power coefficients	c0 = 0.5656	c0 = 0.5656	c0 = 0.3098
	c1= −0.6726	c1= −0.6726	c1= −0.1082
	c2= 1.1477	c2= 1.1477	c2= 0.7972

(2). Heat pump options

The ASHP was re-sized for each climate zone (Table 6) based on the TRNSYS-simulated hourly building loads. The sizing was selected in 1767 W (0.5 ton) increments to meet the following criteria. For cooling-dominated climates (cumulative cooling load > cumulative heating load), the 2^nd-stage cooling capacity was sized between 100 % to 115 % of the maximum cooling load [35]. The capacity was calculated for the outdoor temperature at the maximum cooling condition (rather than the standard rating condition, indoor wet-bulb =19.4 °C, dry-bulb =23.9 °C, outdoor dry-bulb =35 °C [36]). In heating-dominated climates, the 1^st-stage cooling capacity (at the maximum cooling outdoor temperature) was sized to 100 % to 125 % of the maximum cooling load; this allowed for higher heating capacity in the 2^nd stage. The 125 % upper limit avoids excessive cycling and poor dehumidification performance caused by oversized heat pumps [35]. Further, for heating dominated climates, we reduced the heat pump size if a smaller system could meet the heating load and the cooling load using either 1^st or 2^nd stage (with the 100 % to 125 % restriction). This sizing modification only changed the heat pump size in Seattle, reducing it from 8.8 kW to 7.1 kW (2.5 ton to 2.0 ton). The efficiency and ratio of airflow-to-capacity of each heat pump was specified to be the same as the baseline ASHP (Table 3). For sizing the ASHPs, the building loads were calculated with an HRV.

Table 6.

Building loads, heat pump capacities, and borehole sizing parameters

City	MHL (kW)	CHL (kWh)	MCL (kW)	CCL (kWh)	PLFc	PLFh	EFLHc (h)	EFLHh (h)	HP size (ton)	# of bores	Bore length (m)
Miami	6.61	17	8.62	30697	0.61	0.00	3784	14	3.0	6	60
Houston	7.08	1062	8.92	19325	0.43	0.01	2411	280	3.0	4	59
Phoenix	6.98	238	8.41	19431	0.63	0.00	2781	101	3.0	4	64
Atlanta	7.03	3948	6.99	10579	0.52	0.26	1720	808	2.5	3	47
Los Angeles	6.86	227	5.65	6874	0.43	0.05	1625	103	2.0	2	49
Las Vegas	7.22	1132	7.52	13867	0.59	0.23	2173	328	3.0	3	62
San Francisco	7.18	3364	4.24	1902	0.14	0.43	493	1189	2.0	1	42
Baltimore	7.25	9122	7.06	6992	0.41	0.50	1130	1455	3.0	2	46
Albuquerque	7.40	5798	4.92	5840	0.50	0.44	1589	1206	2.0	2	36
Seattle	7.13	10916	4.95	1595	0.02	0.64	380	2569	2.0	2	42
Chicago	8.39	14606	6.30	4531	0.34	0.54	797	1853	3.0	3	51
Boulder	7.79	11075	4.58	2468	0.25	0.41	653	1468	2.0	3	46
Minneapolis	9.64	18768	6.21	3928	0.27	0.60	726	2005	2.5	4	56
Helena	9.51	18652	3.89	1386	0.14	0.50	448	1988	2.0	4	58
Duluth	10.85	27208	5.29	1475	0.00	0.61	382	2608	2.5	6	62

Note: MHL (Maximum heating load); CHL (Cumulative heating load); MCL (Maximum cooling load); CCL (Cumulative cooling load); PLF_c (Part-load factor in cooling design month); PLF_h (Part-load factor in heating design month); EFLH_c (Equivalent full-load hours for cooling); EFLH_h (Equivalent full-load hours for heating).

The GSHP nominal capacity (Table 7) was specified the same as the ASHP (Table 6). A two-stage GSHP unit was modeled after a commercially-available 6.7 kW (2 ton) unit, using the manufacturer’s data for capacity and power at varied entering liquid temperature (ELT), airflow, indoor dry-bulb temperature, and indoor relative humidity. Fig. 7 shows the capacity and power as functions of the ELT (i.e., borehole outlet temperature). For different sizes of GSHP, the efficiency and ratio of airflow-to-capacity of each heat pump was specified as the same as the 6.7 kW (2 ton) system. The GSHP did not have a dedicated dehumidification mode; to provide a fair comparison with the baseline ASHP, this capability was added in the simulation. For dedicated dehumidification, the ratio of dehumidification (i.e., latent) capacity to total cooling capacity is the same as that of the baseline ASHP (Table 3). Similarly, the electrical energy use was scaled by applying the dehumidification-to-cooling ratio from the baseline ASHP. For each of the GSHP simulations, the ventilation system included an HRV.

Table 7.

Components and parameters of TRNSYS models for 6.7 kW (2 ton) GSHP and borehole

Subsystem	Component	TRNSYS type	Main parameters
GSHP	1st stage	Type 919	rated flow rates: 1020 m3/h air, 1.44 m3/h liquid rated capacity: 5273 W total cooling, 4274 W sensible cooling, 3943 W heating rated power: 922 W cooling, 1048 W heating
	2nd stage	Type 919	rated flow rates: 1360 m3/h air, liquid 1. 80 m3/h rated capacity: 6686 W total cooling, 5192 W sensible cooling, 4962 W heating rated power: 1520 W cooling, 1437 W heating
	3rd stage	Type 919	rated flow rates: 1360 m3/h air, 1.80 m3/h liquid rated heating capacity: 4962 W rated heating power: 1437 W supplemental electric resistance heating: 10 000 W
Dedicated Dehumidification for GSHP	1st stage	Type 919	rated flow rates: 985 m3/h air, 1.44 m3/h liquid rated cooling capacity: 5829 W total, 4375 W sensible rated cooling power: 1165 W
	2nd stage	Type 919	rated flow rates: 892 m3/h air, liquid 1.44 m3/h rated cooling capacity: 1564 W total, 410 W sensible rated cooling power: 773 W
Ground heat exchanger	Borehole	Type 557	effective borehole radius: 4.88 cm; borehole spacing: 6.0 m; tube center to center: 6.83 cm tube diameter: 2.54 cm inner, 3.30 cm outer tube conductivity: 1.62 W/(m·K) soil conductivity: 2.43 W/(m·K) soil heat capacity: 2549 kJ/(m3·K) grout (fill) conductivity: 2.617 W/(m·K) fluid density: 980.7 kg/m3 fluid specific heat: 4.396 kJ/(kg·K)

Fig. 7.

Capacity and power of the 6.7 kW (2 ton) GSHP vs. entering liquid temperature

The borehole coupled to the GSHP was sized (Table 6) using the method in Kavanaugh and Rafferty (K&R) [37], which is similar to the method in ASHRAE [38] except that K&R present a more detailed analysis of the long-term temperature penalty; the calculation method is included in the appendix. The maximum and minimum borehole outlet temperatures were specified as 37 °C and −3 °C, and the heat pump efficiency was fixed at the value corresponding to these temperatures (37 °C for cooling, −3 °C for heating). The number of boreholes was adjusted to keep the borehole length shorter than 65 m, staying consistent with the case studies presented in the ASHRAE handbook [38] (in actual installations boreholes may be deeper depending on driller capability and preference). In the simulation, the boreholes were modeled using the duct ground heat storage (DST) model [39]. The DST model inputs (Table 7), including effective thermal conductivity and borehole radius computed from a thermal response test and discussed in [40, 41], are based on the installation at the test house in Gaithersburg.

3.4. Performance indices

The annual heat pump energy consumption (E_HP) was the sum of energy for heating (E_heating), defrost (E_defrost), cooling (E_cooling), dehumidification (E_{dehumidification}), and standby mode (E_standby):

$$E_{HP}=E_{heating}+E_{defrost}+E_{cooling}+E_{dehumidification}+E_{standby}$$ (5)

The annual HVAC energy consumption (E_HVAC) was the sum of energy for the heat pump and ventilation (E_ventilation):

$$E_{HVAC}=E_{HP}+E_{ventilaton}$$ (6)

The annual total building energy consumption (E_building) was the sum of each subsystem consumption, including lighting (E_lighting), plug loads (E_plug), appliances (E_appliance), DHW (E_DHW), and HVAC:

$$E_{building}=E_{lighting}+E_{plug}+E_{appliance}+E_{DHW}+E_{HVAC}$$ (7)

We defined the ratio of HVAC to building energy consumption to evaluate the energy use composition of various HVAC options in different climate zones:

$$R{E}_{HVAC}=\frac{E_{HVAC}}{E_{building}}\times 100\%$$ (8)

We defined the energy saving ratio of the alternative options, as compared to the baseline:

$$\text{ESR}=\frac{E_{baseline}-E_{alterlative}}{E_{baseline}}\times 100\%$$ (9)

The annual net energy (E_net) was the PV generation (E_PV) minus the total building energy consumption:

$$E_{net}=E_{PV}-E_{building}$$ (10)

The PV capacity was sized to meet the net-zero goal in all climate zones (i.e., E_net = 0). We applied a “net-metering” policy (the net electricity was summed annually), so with E_net = 0 there is no energy cost for all the options in all the climates. Therefore, the cost-effectiveness of the options was compared solely on the initial cost. We then used the summed cost of PV and HVAC to compare the various HVAC options, since the cost for other subsystems of the NZEB was not affected by HVAC options:

$$C_{PV+HVAC}=C_{PV}+C_{ventilation}+C_{HP}$$ (11)

where C_PV, C_ventilation, and C_HP are the cost of the PV, ventilation, and heat pump subsystems, respectively.

We used installed cost estimates for the PV and HVAC systems from previous related work [42]; these data were based on averages and contractor estimates for Maryland in 2015. We did not account for local cost variations, but the relative comparisons and equipment selection conclusions are still valid to the degree that heat pump, ventilation, borehole and PV installation costs scale together in different locations. Table 8 lists the installed cost estimates of the PV and HVAC equipment for the NZEB. The installed cost of PV depends on its installed capacity, using a median price of $3.90/W [43]. The installed cost estimates of the HVAC equipment (including ductwork) were taken from previous work [42] and references [44–46]. The median borehole cost was $49.02/m in 2013 [45] and was increased to $53.02/m in 2016 to account for inflation [46]. No tax credits, tax deductions, rebates, or any other financial incentives were applied to the costs. Also excluded are any future costs such as annual maintenance (e.g., replacing air filters, checking performance) and occasional repairs.

Table 8.

Cost of PV and HVAC equipment used in the residential NZEB

Equipment	Cost
PV	$3.90/W
Ventilation without heat recovery	$876
HRV	$4612
ERV	$5203
ASHP (rated 7.1 kW, ~2-ton)	$28 163
GSHP (rated 6.7 kW, ~2-ton)	$36 067
Borehole	$53.02/m

4. Results and discussions

Our analysis focused on PV production, DHW energy, and HVAC energy, since other loads were not dependent on climate zone (lighting, plug loads and appliances were 442 kWh, 2462 kWh and 1898 kWh, respectively). The solar radiation primarily determined the area-specific PV generation of each climate zone (Fig. 8), while the ambient temperature had a small effect on the efficiency. The DHW electricity consumption was most influenced by the mains inlet water temperature (affecting the water heating load) and solar radiation (affecting the SHW capacity), while the difference caused by the indoor temperature (affecting the HPWH efficiency) was small. In general, a lower mains water temperature resulted in a larger DHW electricity use (Fig. 9), excepting in cities with relatively high insolation (e.g., Phoenix, Las Vegas, and Albuquerque).

Fig. 8.

PV generation and solar radiation

Fig. 9.

DHW electricity and mains temperature

4.1. Performance with ventilation options

The ventilation options, seasons, HVAC systems, and climate zones greatly affected the ASHP energy consumption (Fig. 10). Compared to the baseline option without heat recovery, the HRV significantly reduced the energy consumption in cold winter months and hot and dry summer months (e.g., July in Phoenix and Las Vegas). For mild months the HRV provided little benefit because of relatively small positive outdoor-indoor temperature differences in cooling (indoor-outdoor for heating), or even a penalty when there were negative outdoor-indoor temperature differences in cooling (indoor-outdoor for heating). Further, the HRV uses additional electricity for a second fan (ventilation without recovery only uses one fan).

Fig. 10.

Monthly energy consumption of ASHP

The ERV provided similar energy savings as the HRV in cold or hot & dry months, though the ASHP energy was slightly higher because the moisture-porous ERV had a slightly lower sensible effectiveness (Table 5). However, the ERV resulted in significant energy savings over the HRV in the cooling season of humid zones (Miami, Houston, Atlanta, Baltimore, and Chicago) because it reduced the latent ventilation load.

Fig. 11 presents the annual energy consumption of the HVAC, including the ASHP and ventilation. The coldest zones, 6 and 7 (e.g., Minneapolis, Helena, and Duluth), used the most energy due to long and cold heating seasons, while the hottest zones, 1 and 2 (e.g., Miami, Houston, and Phoenix), used the second highest level of energy due to long and hot cooling seasons. Compared to ventilation without recovery, the HRV provided at least some benefit in 12 of the 15 cities, except Miami, Houston, and Los Angeles. The HRV was most beneficial in zones 5 through 7 (e.g., Chicago through Duluth), achieving energy savings of 16.0 % to 20.8 %. The ERV was beneficial in all the climate zones except for Los Angeles, saving energy by 17.3 % to 19.7 % in Chicago through Duluth, the coldest 5 climate zones in Table 4.

Fig. 11.

HVAC energy consumption with different ventilation options

Comparing recovery ventilators, the total energy use with the ERV was less than with the HRV in 8 cities, with savings of at least 5 % in 4 cities: Miami (16.7 %), Houston (16.0 %), Atlanta (9.6 %), and Baltimore (5.5 %). Although the ERV outperformed the HRV by 5.5 % in Los Angeles, both ventilators resulted in more total energy use in this climate than ventilation without heat recovery. The HRV saved more energy than the ERV in 7 cities, but only marginally, with the highest energy savings of 3.1 % in San Francisco. We roughly estimate that differences of less than 2 % were not significant; we don’t have direct measurements comparing HRV vs. ERV performance, but the manufacture’s datasheet lists the uncertainty of effectiveness as 1 %.

Fig. 12 illustrates the energy breakdown of the ASHP without heat recovery (more detailed data in Tables A3 through A5 in the appendix). In the hot and humid zones (e.g., Miami, Houston, Atlanta, and Los Angeles), dehumidification consumed the most energy due to the high-latent and high-cooling loads. These locations were the same as those where the ERV most outperformed the HRV (greater than 5 % total energy savings). In cooler humid zones (e.g., Baltimore, Chicago, and Minneapolis), dehumidification energy was smaller but still accounted for a big fraction of the total; in the remaining dry or cold zones, dehumidification consumed little energy. The 3^rd-stage heating with supplemental electric heating comprised a large percentage in zones 5 through 7 (e.g., Chicago through Duluth), because of the severe ASHP capacity deterioration at low temperatures and the heating-dominant loads. Both the ERV and the HRV yielded significant savings in these zones.

Fig. 12.

Energy breakdown of ASHP without heat recovery

Fig. 13 shows the annual building energy consumption intensity (normalized by floor area excluding the basement and attic) under different ventilation options. These values can be references for the energy consumption levels of low-energy buildings or NZEBs in different climate zones. Fig. 14 depicts the ratio of the HVAC consumption to the total building consumption. When there was no heat recovery, the ratios ranged from 27.5 % to 72.9 % across the different climate zones, with the lowest ratio in San Francisco and the highest ratio in Duluth. The energy ratios were reduced to a range of 23.8 % to 68.7 % by the HRV and to a range of 24.4 % to 69.0 % by the ERV.

Fig. 13.

Annual building energy consumption intensity under different ventilation options

Fig. 14.

Ratio of HVAC to building energy under different ventilation options

Fig. 15 presents the PV capacity and area to meet the net-zero goal under different ventilation options. San Francisco has the smallest installation of 4.69 kW and 21.7 m², with no heat recovery and 4.46 kW and 20.6 m², with the HRV. The largest PV system was in Duluth with 16.53 kW and 76.3 m² without ventilation heat recovery and 14.32 kW and 66.1 m², with the HRV.

Fig. 15.

PV installation for net-zero goal under different ventilation options

The total initial costs of the HVAC and PV systems with different ventilation options are compared in Fig. 16. Only for the coldest climates, e.g., Chicago, Minneapolis, Helena, and Duluth (but not Boulder), the HRV and ERV had lower cost than the baseline option (no heat recovery) due to reduced PV capacity requirements to offset energy consumption. For the other climates, the initial costs of the HRV and ERV were higher by 1 % to 9 %.

Fig. 16.

Initial costs of HVAC and PV under different ventilation options

4.2. Performance with heat pump options

Fig. 17 compares the monthly energy consumption of heat pump options. In the cooling-dominant climate zones with minimal heating loads (Miami, Phoenix, and Los Angeles, as indicated in Fig. 12) and the mild climate zone like San Francisco, the GSHP consumed more energy than the ASHP for most months. For those locations the heat sink temperature of the GSHP was generally higher than that of the ASHP due to high soil temperature (Fig. 18) or short borehole length (Table 6); the heat put into the ground takes time to dissipate and in these cases caused the temperature of the ground immediately surrounding the bore to often be higher than the ambient air temperature. This leads to the interesting result of a GSHP having a lower cooling COP and higher cooling energy consumption, compared to the ASHP. Note that the simulated ground temperature (Fig. 18) is shown for ten years to account for long-term ground temperature change, while the ambient air is only shown for one year since the profile is identical for each year in the TMY2 data. In all the zones except for Los Angeles, San Francisco, and Albuquerque, the GSHP consumes less or about the same amount of energy as the ASHP in the hottest month due to the lower peak heat sink temperature in July (Fig. 18). In climate zones with large heating loads, the GSHP always substantially reduced the energy consumption (relative to the ASHP) in the heating season, owing to the higher heat source temperatures. The ELT of the GSHP was not only much higher but also more stable than the entering air temperature of the ASHP in winter (Fig. 18).

Fig. 17.

Energy consumption of different heat pumps

Fig. 18.

Temperatures of outdoor air, entering liquid, and soil

The total HVAC electricity use was lower with the GSHP than with the ASHP in 11 of the 15 (Fig. 19). The most significant energy savings took place in climate zones 5 through 7 (Chicago through Duluth), reaching 24.3 % to 39.2 %. Though the GSHP was more efficient than the ASHP in Houston and Las Vegas, the energy savings were only 2.7 % and 0.8 %, respectively. In the four cities where the ASHP used less energy than the GSHP, the differences were 13.9 %, 3.5 %, 14.8 %, and 10.0 % in Miami, Phoenix, Los Angeles, and San Francisco, respectively. We roughly estimate that differences of less than 7 % were not significant. Our measurements for the ASHP energy had 5 % uncertainty for the data regression; applying the same 5 % uncertainty to the GSHP manufacturer’s data (and adding the values in quadrature) yields the 7 % threshold.

Fig. 19.

Energy consumption of HVAC with different heat pump options

Compared to the ASHP (Fig. 12), the GSHP (Fig. 20, with more detailed data included in Table A6 in the appendix) significantly reduced the 2^nd- and 3^rd-stage heating (HP and electric), especially for climate zones 6 and 7 (Minneapolis, Helena, and Duluth). In climate zone 5 (Chicago and Boulder), the 3^rd-stage heating was reduced to a negligible fraction. The 3^rd-stage heating was eliminated for climate zones 1 through 4.

Fig. 20.

Energy consumption breakdown of the GSHP

The building energy intensities (Fig. 21) and ratio of HVAC-to-total energy (Fig. 22) were greatly reduced using the GSHP in the cold climates of Chicago through Duluth. Also, using the GSHP significantly reduced the PV system size in these same cold climates by 12.9 % to 23.0 % (Fig. 23).

Fig. 21.

The building energy consumption intensity under different heat pump options

Fig. 22.

The ratio of HVAC to building energy under different heat pump options

Fig. 23.

PV installation for net-zero goal under different heat pump options

The total HVAC+PV initial costs were higher with the GSHP than with the ASHP, by 14 % to 21 % in San Francisco through Duluth and by 25 % to 41 % in Miami through Las Vegas (Fig. 24). Even in the cold climates where the GSHP saved significant energy, the additional cost of the boreholes made the GSHP less cost-effective despite a reduction in PV size needed to reach net-zero energy. However, the boreholes were probably not the most cost-effective size. The ELTs (Fig. 18) often reached neither the minimum nor the maximum specified in the design (Miami, Houston, Atlanta, Seattle, Chicago, Boulder, Minneapolis, Helena, Duluth). In these locations shorter boreholes could be used with only a marginal increase in energy use. We discussed this issue in more detail in the next section (4.3).

Fig. 24.

Initial costs of HVAC and PV under different heat pump options

4.3. Discussion of the borehole sizing methods

The model and assumptions underlying the K&R design method [37] are different from those for the TRNSYS simulations using the DST simulation [39], so the simulated maximum and minimum ELTs did not exactly match values expected from the design (in some cases they differed significantly). For example, the K&R design method is analytical and assumes a cylindrical heat source/sink, one-dimensional radial-only transient conduction, a long-term temperature penalty computed using the average soil temperature change (rather than the temperature change at the bore wall), and a fixed heat pump efficiency. In contrast, the DST implemented in TRNSYS uses a finite difference method to capture both radial and axial conduction (along bore length), implicitly computes the long-term temperature penalty directly at the bore wall, and uses the real-time ELT to determine the heat pump efficiency. We assumed that the additional level of detail in the DST model provided more accurate predictions than the K&R method.

We parametrically varied the number of boreholes in Duluth to better understand the effect on ground and entering liquid temperatures (Fig. 25), energy use (Fig. 26), and initial cost (Fig. 27). The bore length was fixed at 62 m (from Table 6), and the number of bores was decremented from six (design value, Table 6) to two. The minimum ELT was correspondingly reduced from −1.4 °C to −2.1 °C, −3.1 °C, −4.7 °C, and −7.3°C, so the design length based on the ELT target of −3 °C in the TRNSYS simulations was about four boreholes. Interestingly, as the number of boreholes was further reduced to two, the annual energy consumption only increased by 461 kWh, from 9558 kWh to 10 019 kWh. The primary change is an increase in 2^nd-stage heating and a decrease in 1^st-stage heating. The electric resistance energy slightly increased by 101 kWh, from 2435 kWh to 2536 kWh.

Fig. 25.

Effect of number of boreholes on the GHX temperatures

Fig. 26.

Effect of number of boreholes on the HVAC energy consumption

Fig. 27.

Effect of number of boreholes on the initial costs

There are two possible reasons that a large amount of electric resistance heat was needed to supplement insufficient heat pump capacity in Duluth: (1) the heat pump is too small, or (2), the borefield was too small, resulting in low ELT and subsequently degraded heat pump capacity. With fewer boreholes, the change of electric heating is relatively small (Fig. 26). Therefore, it was insufficient heat pump capacity, rather than a small borefield, that required use of the low-efficiency electric resistance heat. Future studies may investigate ways to resize the heat pump to reduce the resistance heating.

The total HVAC+PV cost was the lowest for two boreholes at $101 361, compared to $106 666 for four boreholes and $112 705 for six boreholes (Fig. 27). All the GSHP options had a higher HVAC+PV cost compared to the ASHP, $95 675. Note the GSHP capacity was restricted to be 100 % to 125 % of the maximum cooling load, which is small in Duluth; a larger heat pump would reduce or eliminate the resistance heating and may lower the initial cost. It is possible that the GSHP could achieve a lower cost than the ASHP by increasing GSHP capacity and optimizing the total bore length.

4.4. Discussions and limitations

If the design target is primarily energy savings, rather than cost savings, the equipment selection will be somewhat different. Table A7 summarizes the HVAC options for each climate that might be applied towards the energy-savings goal, where the option listed outperformed one or more options by at least 5 %. It provides suggestions on HVAC system selection for the NZEBs prioritizing energy efficiency rather than initial cost. The table also shows the range of PV areas required to meet the net-zero goal using the various HVAC options.

There are some limitations to this study that need to be considered. The PV system is favored here because we used an optimally-oriented (South) roof unobstructed by trees or other buildings, which is not possible with all houses. We also did not account for snow on PV systems in cold climates, which can significantly reduce their output. Also, some PV areas may be too large to integrate with the roof; for the test house the uppermost south-facing roof, excluding the garage, was 58 m² (total roof area was 116 m²) but the PV area ranged 20.6 m² to 76.3 m² for the 15 cities. We calculated costs everywhere using costs estimates for Maryland, but there will be local cost variations. No applicable energy tax incentives or other credits were applied. Utility rate structures may change the energy cost of the different net-zero configurations; for example, utilities may use time-of-day pricing, or charge/reimburse different amounts for electricity imported/exported to/from the grid (in contrast to the net-metering policy applied here). A more rigorous analysis should consider annual maintenance costs and occasional repairs for the HVAC + PV options. Finally, the conclusions presented here are sensitive to PV price, which has been decreasing in recent years; the average installed cost in Maryland was $3.90/W in 2015 [43] and $3.60/W in 2016 [47]. Reduced PV costs will make energy-saving HVAC technologies relatively less attractive.

In addition, the HRV and ERV options operated continuously, even when there were small positive or negative outdoor-indoor temperature/humidity differences in cooling (indoor-outdoor temperature differences for heating). To avoid ventilation penalties, the control of HRV and ERV could be improved by operating based on the instantaneous outdoor and indoor parameters.

5. Conclusions

This study investigated the energy performance and initial costs of a residential NZEB in different climate zones, using a TRNSYS model validated by a whole year’s operation data. We compared various commercially-available HVAC technologies, including ventilation and heat pump options, in 15 representative cities across the U.S. The HVAC options were each paired with a PV system with capacity to exactly meet the net-zero energy target; this approach highlighted the cost trade-off between NZEB designs with higher efficiency HVAC equipment vs. larger PV arrays. The energy and cost data as well as the PV installation data could help guide HVAC and PV designs for residential NZEBs in different climate zones.

Compared to ventilation without recovery, the HRV saved energy in 12 cities and the ERV saved energy in 14 cities. However, the savings were only cost-effective from an initial-cost perspective in the coldest climates of Chicago, Minneapolis, Helena, and Duluth; the HRV was the most cost-effective option for these climates.

The GSHP used less energy than the ASHP in 11 out of 15 zones across the U.S., particularly in cold climate zones like Chicago through Duluth where savings ranged from 24.3 % to 39.2 %. Despite these energy savings, the GSHP always resulted in higher HVAC+PV first costs when the K&R design method was employed. We explored lowering the GSHP option cost for Duluth by using the target ELT to select four bores rather than six from the K&R design method. Interestingly, even lower costs were achieved with two bores. It is possible that the GSHP option could be cost-competitive with the ASHP in cold climates with optimized total bore length, and perhaps increased GSHP capacity to reduce the electric resistance heating.

To provide suggestions on HVAC system selection for the NZEBs prioritizing energy efficiency, the options outperformed other options by at least 5 % were summarized. For ventilation options, in Phoenix, Los Angeles, and Las Vegas, no heat recovery is an energy-efficient option; in Miami and Houston, the ERV is an energy-efficient option; in Atlanta and Baltimore, both HRV and ERV are energy-efficient options; while in other climates, the HRV is a better option. For heat pump options, in Miami through San Francisco except Atlanta, ASHP is energy-efficient; while in other climates, GSHP is a better option. To provide references to PV selection for the residential NZEBs across the U.S., the required PV installation areas are summarized, being in the range of 20.6 m² to 76.3 m², the lowest in San Francisco and the largest in Duluth.

For the systems studied here, the residential NZEBs were realized at lowest cost when using modest-efficiency HVAC systems paired with larger PV arrays. Future work should consider if cost savings could be achieved with even lower-efficiency HVAC in hot and mild climates or optimized GSHP systems in cold climates. Further, there may be less expensive dehumidification options (e.g., 1^st-stage dehumidification only, and allowing for slight overcooling) suitable for cooler and drier climates. Finally, it would be useful to study if the strategy of using larger PV arrays is still valid for orientations that are not optimal (e.g., facing east or west) or partially shaded.

Nomenclature

C

cost, $

C_f

diffusion factor

c_p

specific heat, kJ∕(kg·°C)

E

energy consumption, kWh

F

heat loss factor

G

solar radiation, kJ/(h·m²)

L

length, m

m

mass flow rate, kg/s

N

number of boreholes

P

power, W

Q

thermal energy, kWh

q

heat rate, W

R

thermal resistance, (m·K)/W

RE

ratio of energy

S

bore separation distance, m

r

radius, m

t

temperature, °C

W

width, m

η

efficiency

ε

effectiveness

ρ

density, kg/m³

Abbreviations

ACCA

Air Conditioning Contractors of America

ASHP

air-source heat pump

ASHRAE

American Society of Heating, Refrigerating, and Air-Conditioning Engineers

CCL

cumulative cooling load

CHL

cumulative heating load

COP

coefficient of performance

DHW

domestic hot water

DST

duct ground heat storage model, for borehole ground heat exchangers

EFLH

equivalent full-load hours

ELT

GSHP entering liquid temperature (i.e., borehole outlet temperature)

ERV

energy recovery ventilator

ESR

energy saving ratio

GSHP

ground-source heat pump

HP

heat pump

HPWH

heat pump water heater

HRV

heat recovery ventilator

HVAC

heating, ventilation, and air-conditioning

K&R

Kavanaugh and Rafferty design method [37]

MCL

maximum cooling load

MHL

maximum heating load

NIST

National Institute of Standards and Technology

NZEB

net-zero energy building

PLF

part-load factor

PV

photovoltaic

SHW

solar hot water

TMY

typical meteorological year

TRNSYS

Transient System Simulation software tool

Appendix

(1). TRNSYS model of the residential NZEB [31, 32]

-Thermal zone division

The NZEB 3-D model, shown in Fig. A1, was created using the as-built architectural plans. An in-built Type 56 in TRNSYS was used to model the building, and the layers of material used in the walls, roof, floors, and ceilings were specified in accordance with the plans. The model was divided into four zones: the basement (Zone 1), the main/first floor (Zone 2), the second floor (Zone 3), and the attic (Zone 4). Each zone was assumed to have a uniform temperature and humidity.

The design of this NZEB can meet the requirement of ultra-efficient houses across most U.S. climate zones, per one of the publications on this NZEB [48]. To make the study focused on PV and HVAC, the same building design that can meet the ultra-efficient requirement in most climates was adopted.

Fig. A1.

Thermal zone division of the NZEB in TRNSYS

-Heating and cooling season

The heat pump thermostat transitioned to heating mode when the temperature dropped1.67 °C below the cooling setpoint and switched to cooling mode when the temperature rose1.67 °C above the heating setpoint. This control logic was programmed in the TRNSYS simulation by creating a new, simple type that read the zone temperature, heating setpoint, and cooling setpoint each time step and determined which mode the heat pump was operating in.

-Occupancy schedule

The schedules for the four virtual occupants were based on the Building America profiles [23]. Each occupant was assumed to generate a constant 70 W of sensible heat and 45 W of latent heat, which was the adjusted load per person for seated and very light work. The occupants’ sensible load was emulated using resistance heaters situated on the first and second floors. The occupants’ latent load was converted from 45 W to 0.07 L/hr of water vapor input, using the latent heat of vaporization. The total volume of water generated by the occupants each day was emulated in the house using ultrasonic humidifiers.

-Plug Load Schedule

The electric plug loads were assumed to be used or owned by at least 50% of households [23]. The electric plug loads followed a weekly schedule. Each plug load was represented either by the actual appliance or by an electric resistance heater box. The sensible and latent load fractions for the actual appliances were assumed to be 0.734 and 0.20, respectively.

-Lighting Schedule

The lighting schedule was based on the movements of the virtual occupants, turning on when an occupant entered a room and turning off when the occupant left the room. Because the basement and attic were unoccupied, only lights on the first and second floors were activated according to the weekly schedule. All of the lighting contributes to the sensible load.

-Appliance Schedule

The appliances included a washing machine, clothes dryer, dishwasher, oven, cooktop range, refrigerator, and a microwave, all of which were located on the first floor. The appliance power was both simulated and measured separately from the plug load. The sensible and latent load fractions differed for each appliance. The latent load generated during typical daily cooking events was estimated and this equivalent volume of water was released in the NZEB each day by ultrasonic humidifiers.

-Moisture Schedule

The moisture generated by a family of four’s occupancy and cooking was emulated using ultrasonic humidifiers. The volume of water varied from day to day depending on the occupancy schedule and the number of cooking events. Table A1 lists the daily volume of water introduced to through humidifiers located in the kitchen. This volume of water was introduced to Zone 2 in the TRNSYS model at a constant rate during hours in which the house was occupied each day.

Table A1.

Daily moisture generated by cooking events and occupants

Day	Moisture from cooking (liters/gallons)	Moisture from people (liters/gallons)	Total moisture (liters/gallons)
Monday	0.99/0.26	5.83/1.54	6.83/1.80
Tuesday	0.74/0.20	3.87/1.02	4.61/1.22
Wednesday	0.74/0.20	4.00/1.06	4.74/1.25
Thursday	0.74/0.20	3.87/1.02	4.61/1.22
Friday	0.74/0.20	4.00/1.06	4.74/1.25
Saturday	0.74/0.20	3.87/1.02	4.61/1.22
Sunday	0.99/0.26	5.63/1.49	6.22/1.75

-Water Draw Schedule

There are five types of water draws: sink draws, baths, showers, clothes washer cycles, and dishwasher cycles. The showers, the baths, and some of the sink water draw events occurred in the master bedroom on the second floor. The other sink events occurred in the kitchen on the first floor. The total water volume used in a week was approximately 2,229 L (589 gal). The sensible and latent heat gain caused be each type of event, as well as other details about each type of event, are listed in Table A2.

Table A2.

Water draw events

	Water draw per event (L)	Length of water draw (min)	Number of events per week	Water temperature (°C)	Sensible gain (kJ/event)	Latent gain (kg/event)
Sinks	1.78	1	280	40.56	12.28	0.002
Bath	116.55	11	2	43.33	1713.2	0
Short shower	39.42	10	21	43.33	581.05	0.226
Long shower	59.13	15	5	43.33	871.58	0.367
Clothes washer	56.7	5	6	48.9	-	-
Dishwasher	7	3	5	48.9	-	-

(2). The borehole sizing method

The Kavanaugh and Rafferty (K&R) method for borehole sizing is briefly introduced here; more details for the determination of each parameter or term can be found in [37].

The required length for cooling (L_c, m) and heating (L_h, m) is respectively calculated by:

$$L_c=\frac{q_a{R}_{ga}+q_{cond}\left(R_b+PL{F}_m{R}_{gm}+F_{SC}{R}_{gst}\right)}{t_g-\frac{t_{wi}+t_{wo}}{2}-t_p}$$ (A1)

$$L_h=\frac{q_a{R}_{ga}+q_{evap}\left(R_b+PL{F}_m{R}_{gm}+F_{SC}{R}_{gst}\right)}{t_g-\frac{t_{wi}+t_{wo}}{2}-t_p}$$ (A2)

where F_SC is the short-circuit heat loss factor; PLF_m is the part-load factor during design month; q_a is the net annual average heat transfer to ground (W); q_evap is the heat pump evaporator heat extraction rate from ground (W); q_cond is the heat pump condenser heat rejection rate to ground (W); R_ga, R_gm, and R_gst are the effective thermal resistances of ground with an annual pulse, monthly pulse, and peak short term pulse (1 h to 6 h recommended) ((m·K)/W); R_b is the thermal resistance of bore ((m·K)/W); t_g is the undisturbed ground temperature (°C); t_p is the temperature penalty for interference of adjacent bores (°C); t_wi and t_wo are liquid temperature at heat pump inlet and outlet (°C). Note that the heat transfer rate, building loads, and temperature penalties are positive for heating and negative for cooling. The annual heat transfer q_a is determined by:

$$q_a=\frac{q_{cond}\times EFL{H}_c+q_{evap}\times EFL{H}_h}{8760}$$ (A3)

where EFLH_c and EFLH_h are the equivalent full-load hours in cooling and heating modes.

The temperature penalty for an internal bore is computed by dividing the stored energy by the heat capacity of the soil within the of the rectangular prism symmetry boundary:

$$t_{p,in}=\frac{Q_{stored}}{\rho c_p{S}_{bore}^{2}L}$$ (A4)

where Q_stored is the energy stored (or extracted from) within adiabatic symmetry boundary (kWh); ρc_p is the soil volumetric capacity (kJ/(m³·°C)); S_bore is the bore separation distance (m); L is the total bore length (m).

The energy that would have otherwise been stored in hollow cylinders of soil beyond the symmetry boundary is:

$$Q_{\text{stored}}=\sum _{r=S_{\mathrm{bore}}/2}^{r_{\max }}\rho c_p\pi L\left(r_o^2-r_i^2\right)\times \left(\text{Δ}{t}_r\right)$$ (A5)

where r_o and r_i are the outer and inner radius of hollow soil cylinder (m); r_max is the maximum considered radius; Δt_r is the soil temperature change at average radius r = (r_o + r_i)/2, (°C).

Finally, t_p is calculated by prorating t_p,in based on the number of bores with a particular adjacency: interior, side, corner, midrow, and end, as well as accounting for heat diffusion at the bottom of the borefield:

$$t_p=\frac{N_{int}+0.75N_{side}+0.5N_{corner}+0.5N_{midrow}+0.25N_{end}}{\left(N_{int}+N_{side}+N_{corner}+N_{midrow}+N_{end}\right)\times C_{fHoriz}}\times t_{p,in}$$ (A6)

where N_int, N_side, N_corner, N_midrow, and N_end are respectively the number of interior, side, corner, middle-of-row, and end boreholes; C_fHoriz is the bottom diffusion factor, which is the ratio of surface area of the sides and bottom of the borefield to the surface area of the sides:

$$C_{fHoriz}=\frac{\left[2\times L_{bore}\times \left(W_{field}+L_{field}\right)\right]+W_{field}\times L_{field}}{2\times L_{bore}\times \left(W_{field}+L_{field}\right)}\times t_{p,in}$$ (A7)

where L_bore is the individual bore length (the total bore length divided by the number of bores). The borefield length (L_field) and width (W_field) are:

$$L_{field}=S_{bore}\left(N_{long}-1\right)$$ (A8)

$$W_{field}=S_{bore}\left(N_{wide}-1\right)$$ (A9)

where the borefield has N_long bores in the length direction and N_wide bores in the width direction.

(3). Energy breakdown data

Table A3 to A6 provide the detailed energy breakdown data of the ASHP (without heat recovery, with HRV, and with ERV) and the GSHP (with HRV).

Table A3.

Energy breakdown of ASHP without heat recovery (kWh)

Climate zone	1st -stage heating	2nd -stage heating	3rd -stage HP heating	3rd -stage electric heating	Defrost	1st -stage cooling	2nd -stage cooling	Dehumidification	Standby
Miami, FL	2.7	0.0	0.0	0.0	0.0	3553.0	499.1	5051.8	180.2
Houston, TX	442.2	66.6	5.3	16.5	38.6	1850.4	286.0	4055.4	247.5
Phoenix, AZ	143.0	2.6	3.4	10.8	5.1	4135.7	1446.2	107.5	310.8
Atlanta, GA	1090.9	503.8	16.4	67.3	133.3	975.5	228.9	1914.7	284.5
Los Angeles, CA	141.1	2.9	2.6	12.0	0.0	345.1	100.4	1250.3	359.0
Las Vegas, NV	575.7	33.8	1.7	5.8	31.3	3263.8	730.3	6.6	333.5
San Francisco, CA	1034.2	135.1	1.3	6.0	3.0	25.8	11.4	461.4	379.1
Baltimore, MD	2382.4	1060.4	54.4	195.2	304.3	499.6	23.0	1592.9	294.4
Albuquerque, NM	985.2	1301.8	60.2	302.0	309.2	879.9	417.6	88.6	317.3
Seattle, WA	2214.7	1275.5	6.6	29.3	73.6	34.2	18.4	253.8	312.8
Chicago, IL	2930.7	2158.8	281.6	1068.0	555.4	205.7	6.7	1254.9	274.4
Boulder, CO	1346.3	2121.7	332.2	1801.8	552.6	314.2	76.7	145.5	302.0
Minneapolis, MN	2023.0	2889.9	919.6	4999.3	874.7	119.7	32.4	1110.7	242.4
Helena, MT	1800.9	2898.9	703.0	4650.7	783.8	150.6	27.9	42.8	262.6
Duluth, MN	3347.2	3504.5	1354.9	7179.7	1115.4	5.6	11.5	153.7	246.6

Table A4.

Energy breakdown of ASHP with HRV (kWh)

Climate zone	1st -stage heating	2nd -stage heating	3rd -stage HP heating	3rd -stage electric heating	Defrost	1st -stage cooling	2nd -stage cooling	Dehumidification	Standby
Miami, FL	0.4	0.0	0.0	0.0	0.0	3633.2	401.2	5040.3	179.7
Houston, TX	253.8	24.4	1.7	5.5	25.7	1917.6	203.3	4114.0	250.3
Phoenix, AZ	40.6	0.6	1.8	5.8	3.1	4317.7	1071.2	98.8	313.8
Atlanta, GA	783.3	279.9	4.5	19.5	107.7	1078.5	207.5	1826.0	298.6
Los Angeles, CA	32.0	0.2	1.2	5.7	0.0	549.7	151.4	1059.3	361.9
Las Vegas, NV	274.9	6.4	1.8	6.0	20.7	3359.9	507.4	4.4	341.7
San Francisco, CA	622.0	45.5	1.3	6.0	2.6	77.8	21.3	260.6	403.0
Baltimore, MD	1943.2	635.3	17.9	64.5	257.6	535.7	20.8	1701.6	308.7
Albuquerque, NM	757.3	805.2	14.5	71.0	250.9	1006.8	403.4	70.4	334.8
Seattle, WA	1881.2	710.9	1.4	6.7	64.0	78.7	14.9	281.0	336.4
Chicago, IL	2621.7	1457.6	159.4	621.7	481.7	259.5	5.9	1187.9	296.8
Boulder, CO	1130.1	1559.1	183.7	1018.7	479.2	389.9	75.3	119.2	325.6
Minneapolis, MN	1873.6	2349.1	588.4	3422.5	777.9	152.3	21.7	1159.4	262.6
Helena, MT	1652.7	2277.7	447.6	3176.3	705.0	208.1	32.4	43.9	288.2
Duluth, MN	3075.1	2837.3	926.4	5250.0	996.7	7.6	10.7	181.9	272.9

Table A5.

Energy breakdown of ASHP with ERV (kWh)

Climate zone	1st -stage heating	2nd -stage heating	3rd -stage HP heating	3rd -stage electric heating	Defrost	1st -stage cooling	2nd -stage cooling	Dehumidification	Standby
Miami, FL	0.4	0.0	0.0	0.0	0.0	3931.8	375.0	3058.3	235.0
Houston, TX	260.3	25.5	1.8	6.0	26.6	2079.5	188.6	2719.4	291.6
Phoenix, AZ	44.2	0.7	1.7	5.7	3.1	4341.6	1034.4	38.7	315.5
Atlanta, GA	800.4	290.1	5.7	22.8	109.5	1160.1	197.1	1170.4	322.7
Los Angeles, CA	35.8	0.2	2.1	9.3	0.0	560.8	140.8	854.3	373.9
Las Vegas, NV	288.8	7.6	1.7	5.7	21.5	3358.8	506.8	1.4	341.5
San Francisco, CA	642.4	48.2	1.2	5.5	2.6	71.0	20.3	263.5	403.4
Baltimore, MD	1967.4	658.9	19.3	69.3	260.5	577.1	17.4	1223.2	322.8
Albuquerque, NM	767.4	832.0	15.6	80.3	252.1	1018.3	384.4	19.3	335.9
Seattle, WA	1901.5	736.8	1.4	6.7	64.0	72.9	13.5	248.6	338.1
Chicago, IL	2641.4	1496.7	163.5	639.8	485.9	291.1	6.4	876.2	305.0
Boulder, CO	1139.7	1594.6	189.9	1051.7	484.3	391.9	74.5	63.4	327.1
Minneapolis, MN	1884.0	2382.4	609.3	3493.5	784.1	147.3	19.2	914.5	273.1
Helena, MT	1659.4	2314.1	458.5	3241.2	710.3	210.0	27.6	28.3	287.5
Duluth, MN	3089.6	2881.8	938.8	5339.0	1003.4	8.6	10.4	152.8	272.4

Table A6.

Energy breakdown of GSHP with HRV (kWh)

Climate zone	1st -stage heating	2nd -stage heating	3rd -stage HP heating	3rd -stage electric heating	Defrost	1st -stage cooling	2nd -stage cooling	Dehumidification	Standby
Miami, FL	0.0	0.0	0.0	0.0	0.0	2334.9	448.9	7685.6	122.8
Houston, TX	184.4	2.8	0.0	0.0	0.0	1021.7	141.1	5017.1	224.9
Phoenix, AZ	24.7	2.6	0.0	0.0	0.0	4049.5	1273.9	415.1	296.9
Atlanta, GA	804.8	3.3	0.0	0.0	0.0	625.4	124.0	2270.9	281.4
Los Angeles, CA	15.2	3.5	0.0	0.0	0.0	524.0	193.9	1472.1	327.3
Las Vegas, NV	194.0	2.9	0.0	0.0	0.0	3188.4	725.7	31.0	333.4
San Francisco, CA	698.5	3.7	0.0	0.0	0.0	120.0	39.6	377.9	380.3
Baltimore, MD	2224.3	5.2	0.0	0.0	0.0	248.1	14.4	1732.1	296.8
Albuquerque, NM	1308.8	4.8	0.0	0.0	0.0	932.9	548.4	134.0	321.9
Seattle, WA	2589.3	4.0	0.0	0.0	0.0	69.7	18.1	274.2	307.3
Chicago, IL	3501.2	38.7	39.8	176.6	0.0	96.7	4.1	1105.5	285.7
Boulder, CO	2323.7	161.2	19.7	128.6	0.0	285.6	51.1	139.7	308.1
Minneapolis, MN	3762.8	288.1	226.6	1195.7	0.0	50.5	9.2	960.4	237.8
Helena, MT	3241.6	255.0	161.7	1013.3	0.0	155.9	22.1	62.8	276.6
Duluth, MN	4615.0	1110.2	446.1	2382.1	0.0	1.8	6.4	128.0	228.5

(4). Energy-efficient HVAC options and required PV areas

Table A7 presents the summarized “energy-efficient” HVAC options with energy savings of more than 5 %. For ventilation options, the HRV is listed if it saved more than 5 % compared to no heat recovery. Both HRV and ERV are in the table if the HRV saved more than 5 % compared to no heat recovery, and the ERV saved more than 5 % compared to the HRV. Just the ERV is listed if it saved more than 5 % compared to no heat recovery, but the HRV did not. For heat pump options, the GSHP is listed if it saved more than 5 % compared to the ASHP.

Based on Fig. 15 and Fig. 23, the minimum and maximum PV areas required to meet the net-zero goal with the investigated HVAC options were also provided (Table A7). These values provide references to PV selection for the residential NZEBs across the U.S.

Table A7.

Energy-efficient HVAC options and required PV areas in different climate zones

Climate zone	Representative city	Energy-efficient ventilation option	Energy-efficient heat pump option	PV installation area (m2) *
				Min	Max
1A	Miami, FL	ERV	ASHP	38.4 (ERV)	46.8 (GSHP)
2A	Houston, TX	ERV	ASHP	36.6 (ERV)	40.1 (no HR**)
2B	Phoenix, AZ	no heat recovery	ASHP	27.0 (ERV)	27.8 (GSHP)
3A	Atlanta, GA	HRV, ERV	GSHP	29.3 (ERV)	31.8 (no HR)
3B	Los Angeles, CA	no heat recovery	ASHP	21.2 (no HR**)	22.8 (GSHP)
3B	Las Vegas, NV	no heat recovery	ASHP	24.3 (HRV)	24.8 (no HR)
3C	San Francisco, CA	HRV	ASHP	20.6 (HRV)	21.7 (no HR)
4A	Baltimore, MD	HRV, ERV	GSHP	34.7 (GSHP)	39.7 (no HR)
4B	Albuquerque, NM	HRV	GSHP	22.3 (GSHP)	24.9 (no HR)
4C	Seattle, WA	HRV	GSHP	38.3 (GSHP)	40.8 (no HR)
5A	Chicago, IL	HRV	GSHP	38.5 (GSHP)	48.8 (no HR)
5B	Boulder, CO	HRV	GSHP	26.6 (GSHP)	35.3 (no HR)
6A	Minneapolis, MN	HRV	GSHP	41.3 (GSHP)	60.2 (no HR)
6B	Helena, MT	HRV	GSHP	36.9 (GSHP)	54.9 (no HR)
7	Duluth, MN	HRV	GSHP	51.3 (GSHP)	76.3 (no HR)

^*The PV area required to meet the net-zero goal (E_net=0) under the investigated HVAC options.

^**No HR = No heat recovery