Laboratory and Field Evaluation of a Gas Heat Pump-Driven Residential Combination Space and Water Heating System.
Representing a combined 4.6 quads per year, the natural gas and propane consumed to heat and provide hot water to U.S. homes is a significant energy expenditure and, as a result, source of greenhouse gas (GHG) emissions. The majority of this expenditure and resulting emissions are from single family and multifamily buildings with fewer than five units, a combined 90.3 million housing units in the U.S. These homes, the 57% that are heated and 56% that generate domestic hot water (DHW) with natural gas or propane, are predominantly served by low-efficiency equipment (EIA RECS, 2009/2015). Concerning gas-fired water heaters, which are historically 10% tankless and 90% storage products sold each year (Energy Star, 2010), only 5% of the storage water heaters are 0.67 Uniform Energy Factor (UEF) or greater, qualifying for EnergyStar with the rest near or at the minimum allowable efficiency (Ryan, 2016). For central warm-air furnaces, which are in 85% of homes heated with natural gas or propane, less than half have a rated efficiency of 90% or greater, as an Annual Fuel Utilization Efficiency (AFUE) (DOE SNOPR, 2016). Higher efficiency gas-fired space and water heating equipment have struggled historically for a number of reasons, including:
* Low, stable utility costs: Since 2009, depressed natural gas prices have increased the financial payback periods of all high-efficiency heating equipment. This diminishes the incentive to upgrade from minimum-efficiency equipment. This is not the case for propane, which has increased by 2.5X over a 20-year period. However, propane is a niche fuel, for every home fueled by propane there are 11-13 homes fueled by natural gas.
* Infrastructure requirements: With improved efficiency, equipment costs are expected to be higher, a 49% incremental cost for 98% over 80% AFUE furnaces (DOE SNOPR, 2016) and a 170% incremental cost for 0.77 over 0.62 UEF storage water heaters (Glanville, 2016). However, installation costs are important, including: increasing the size/capacity of gas piping, conversion of venting, addition of combustion condensate handling/neutralizing equipment, and often for storage water heater upgrades, adding electrical service. These combined costs can often eliminate net cost savings for high-efficiency gas-fired equipment.
* Reduced heating and DHW loads with time: Through improvements in building envelopes, including insulation and windows, and internal migration from cold to warm climate regions, the energy consumption for heating homes has steadily declined for several decades. Historically the largest energy expenditure, the U.S. Department of Energy (DOE) has noted that since 2009, energy inputs for heating and cooling are no longer the majority (EIA RECS, 2009). Coupled with increased efficiency of equipment, overall the energy use intensity (EUI) of residential buildings has had an overall decline, 37.4% from 1980 to 2009 (Nadel, 2015). Similar reductions of DHW loads were observed due to more water-efficient fixtures and a steady drop in home occupancy, with single-occupant homes at an all-time high (U.S. Census, 2010). Manufacturers have responded through reducing equipment sizes, however these declining loads have an overall effect of extending the payback period of higher efficiency equipment.
One method of addressing these challenges while improving the cost-effectiveness of high-efficiency gas heating equipment are combination space and water heating systems ("combi" systems). Combi systems offer advantages through coupling space and water heating with a single thermal engine. By replacing two components (e.g. furnace and storage water heater), the incremental equipment cost and installation costs can be offset by a greater delivered efficiency of space heating and DHW. As a result for a given home, the unacceptably long payback periods of upgrading from an 80% to [greater than or equal to] 90% AFUE furnace and, more commonly, any water heating product above the allowable minimum UEF can become attractive. While additional components are often necessary, such as circulation pumps and a hydronic air handling unit (AHU), those costs saved by "bundling" equipment and infrastructure upgrades outweigh these system costs with as-installed system efficiencies of 90% or greater (Kingston, 2016). For packaged combi systems and field-engineered systems alike, simple improvements in controls can further improve cost-effectiveness with overall payback periods in "low load" homes as short as 2.5 years (Schoenbauer, 2016). While hydronic heating distribution systems (delivering heat via radiators, emitters, etc.) are more common abroad, the vast majority of U.S. homes have forced-air heating distribution, with 73% of single family homes with central A/C and of those with gas heating, 88% have central warm-air furnaces. As a result, the focus of this study is combi systems with forced-air distribution.
Concerning the taxonomy of combi systems in forced-air heating applications, combi systems are generally deployed in one of two manners, using potable and non-potable heaters, in open and closed arrangements respectively:
* Potable (Open): Using either a storage or tankless water heater, the heater is plumbed to a hydronic AHU in a recirculation loop and, often using a thermostatic mixing valve, DHW is drawn off of the loop at a desired temperature. As the heater is directly providing DHW to the home, a popular heater option is a tankless water heater, with the advantage of high turndown ratios (up to 20:1). This is advantageous as space heating and DHW loads are quite different, the former seasonal, steady, and relatively low magnitude (commonly 20-40 kBtu/h or 5.9-11.8 kW) and the latter year-round, highly intermittent, and frequently large magnitude (commonly peaking at 100-180 kBtu/h or 29.3-52.3 kW). Tankless water heaters can cover this range, capable of ramping up/down to meet intermittent DHW loads, while potentially eliminating the need for storage. By contrast, combi systems based on storage water heaters can have issues with output capacity during periods of high DHW and/or space heating loading (Kingston, 2016). Based on tankless or storage water heaters, these combi systems are "open" in the sense that space heating is provided by heated potable water, which can be drawn off as DHW. As a result, care must be taken with respect to water quality and water-borne bacteria (e.g. Legionella). Also, these combi systems generally have a maximum operating temperature for space heating of 140[degrees]F (60[degrees]C). In general, codes do not permit the use of water heaters in closed, heating-only, applications and may place special requirements on their use in the aforementioned "open" system arrangement.
* Non-potable (Closed): Using a boiler as the central heating plant, directly tied to the hydronic AHU, these non-potable central heating plants have advantages over open systems including reduced system complexity and likelihood for system short-cycling in space heating modes, ability to operate with higher space heating temperatures (generally up to 180[degrees]F/82.2[degrees]C), broad acceptance by building codes, and minimal concerns with water quality/bacteria. Generally, in this arrangement the boiler will be tied to an indirect storage tank in parallel to the AHU to provide potable DHW, which this buffer storage can eliminate temporary loss of DHW common with tankless water heater-based combi systems caused by system cycling/adjustments ("cold water sandwich" or "cold slug"). However, boiler manufacturers have introduced compact "combi" boilers which can independently provide space heating and DHW. Based on conventional boilers, with indirect storage tanks, or "combi" boilers, these combi systems tend to have greater equipment and installation costs over potable/open combi systems. This primarily concerns the heater itself, where gas boilers can have 2-3X the equipment cost of a tankless water heater.
With advantages and disadvantages between potable and non-potable based combi systems, depending on specifics of the home and retrofit versus new construction scenarios, collectively combi systems are an emerging solution to improve the energy efficiency of residential space and water heating systems. While the industry continues to improve cost-effectiveness and efficiency, through development of packaged systems, design guidance to minimize return water temperatures for consistent high-efficiency operation, and controls to limit short cycling, these improvements are incremental in nature. Like standalone gas-fired space and water heating equipment, these combi systems are limited to upper "condensing" level efficiency, an effective 98% upper limit. Beyond this limit, manufacturers and researchers have developed and demonstrated combi systems driven by heat pumps, both electrically-driven (Eklund, 2016) and gas-fired (Toppi, 2014 and Garrabrant, 2016). As an option for the previously discussed homes with gas heating, a gas heat pump-driven combi system may offer significant energy savings for delivered space and water heating over baseline and conventional "condensing" efficiency options. Building on these efforts, in this paper, the authors outline the application of a recently developed low-cost gas absorption heat pump (GAHP) to residential combi systems. In addition to the benefits of higher operating efficiency, up to 45% therm savings over baseline, and as part of a non-potable (closed) system this GAHP has the following added benefits: it uses a natural refrigerant/working pair and combustion occurs outdoors, eliminating IAQ/venting concerns.
Low-Cost Gas Absorption Heat Pump Combi System
In the pursuit of a cost-competitive GAHP intended for residential retrofits, hydronic or forced-air heating distribution systems, the authors developed and demonstrated several prototype air-source GAHPs through laboratory and early-stage field evaluations with support from a boiler and water heater manufacturer, as described previously (Glanville, 2017). Through selection of the single-effect ammonia-water vapor absorption cycle in a heating-only mode and a focus on easily manufactured components, the team sought to demonstrate an economical GAHP, at a projected equipment price of less than $5,000, 30-50% that of GAHPs currently available. As an absorption heat pump, lifting the pressure of the liquid refrigerant/absorbent solution is performed by a solution pump, however the primary input is thermal energy from the multi-stage, 55,000 Btu/hr (16.1 kW) gas burner required to drive the refrigerant vapor from its absorbed state in the desorber (or "generator"). The air-to-brine GAHP is installed wholly outdoors and has a nominal output of 80,000 Btu/hr (23.4 kW) with 4:1 system modulation, and is without backup or supplemental heating.
Outside of this "thermal compressor", comprised of an absorber, desorber, solution pump, solution heat exchanger, and rectifier, the balance of the heat pump components are familiar, the condenser, expansion valve, evaporator, and refrigerant heat exchanger. The energy input from gas combustion drives the vapor absorption cycle, which generates a 'refrigeration effect', from which approximately 40% of the GAHP heat output to the hydronic loop is drawn from the evaporator to the condenser. The remaining 60% of heat output is heat recovery from (a) the hot weak solution and its heat of sorption within the absorber and (b) the warm flue gases exiting the direct-fired desorber at ~300[degrees]F (~150[degrees]C) via a flue-to-hydronic condensing heat exchanger (CHX). While operating, the early generation prototypes's peak power draw was up to 600 W, needed to drive the evaporator fan, solution pump, combustion blower, and controls. As a heat pump, the net heating output exceeds the energy input with operating coefficients of performance (COP) ranging from 1.4 to 1.9 and by virtue of the CHX, it operates in the "condensing mode".
Two first-generation prototype GAHPs were designed, built, and laboratory-tested in a research project funded by the U.S. Department of Energy (DOE) with support from gas utilities. At the close of this effort, the team demonstrated operating cycle COPs of 1.65 (HHV Basis) at a 100[degrees]F (37.7[degrees]C) hydronic loop return and 47[degrees]F (8.3[degrees]C) outdoor condition, with lower efficiencies recorded at higher return temperatures and/or colder ambient temperatures, down to a cycle COP of 1.2 (HHV Basis) at -13[degrees]F (-25[degrees]C) with a hydronic return temperature of 95[degrees]F (35[degrees]C). Through testing to the ANSI Z21.40.4 method, an AFUE of 139% for U.S. Climate Region IV was demonstrated for the 2nd gen. prototype. On emissions, during this laboratory test program the GAHP had Ultra-Low NOx combustion, below 14 ng NOx/J output. A full review of this GHP development is provided in reporting to the U.S. DOE (Garrabrant, 2015). This paper is an overview of a) recent field testing of the GAHP prototype applied as a combi system and b) companion laboratory testing simulating combi system operation (Figure 1).
FIELD EVALUATION OF GAHP COMBI SYSTEM
Following the aforementioned laboratory development and testing of early-generation prototypes, two units were installed at field sites in close proximity to the participating manufacturer: one providing space heating to a commercial warehouse (previously described in Glanville, 2017) and the other providing combined water and space heating to a single-family residence, the topic of this paper. The residence is a 1,800 sf (167 [m.sup.2]) single-family home in Tennessee (Mixed-Humid climate) with 3-4 occupants, with a college-aged occupant with intermittent occupancy. The GAHP combi system was monitored over a period of 15 months, beginning in the middle of the 2015-16 heating season and completing at the end of the 2016-17 heating season. As a result, there were three distinct phases: 1) a 1st gen. prototype ("Alpha 3") GAHP provided space/water heating to the residence through the duration of the first heating season, 2) the same unit provided DHW-only until the start of the second, during which the manufacturer elected to replace the indirect storage tank to improve system delivered efficiencies, and 3) upon the initiation of the 2016-17 heating season the "Alpha 3" was replaced by a 2nd gen. ("Beta 1") GAHP and operated over the second heating season. While small improvements and changes were made throughout, a substantial change to system controls was implemented half-way through the second heating season, which in total sought to limit short-cycling, that is to reduce the number of GAHP on-cycles while lengthening space heating/DHW/combined cycles. Overall, this GAHP operated for 1,783 cycles for a total of 1,344 hours, 31% space heating (SH) cycles, 52% DHW-only cycles, and 17% combined DHW/SH cycles.
As a disclaimer, this early field evaluation differs from that of commercially-available products due to the early stage of this technology. Results are more for late-stage product development than are representative of a commercialized product, however performance extrapolations are made nonetheless. The purpose was to:
* Obtain operating hours in an uncontrolled outdoor environment and tied to actual heating loads to provide early feedback regarding design, control and operational weaknesses.
* Complete the process of specifying and installing the entire heating system (hydronic loop, pumps, air-handlers, indirect storage tank, pumps, system controls, condensate management, gas line, power, etc.), to develop best practices.
Continuous data acquisition was performed by a standalone data acquisition system (DAS) with data downloaded on a weekly basis from the instruments outlined in Figure 2. Continuous data collection included: representative indoor/outdoor temperatures, natural gas and power inputs, critical recirculating loop temperatures and flow rates, DHW characteristics (hot/cold water, flow, and tank temperatures), supply/return air temperatures at AHU with fan speed setting (Low/Mid/High), critical measurements within and external to the GAHP. Data were taken every 15 seconds if the GAHP was firing or if a DHW draw was occurring, otherwise data were taken every 5 minutes. Third parties and batch measurements were used to quantify: natural gas HHV, combustion excess aeration, manifold fuel pressure, storage tank volume, and fraction of glycol/water in GAHP loop. Data collected were used to quantify space and water heating loads, GAHP energy consumption, and GAHP output and efficiency. Concerning the final grouping, note that the GAHP output is calculated as heat delivered to the recirculating loop with the efficiency defined as COP on a gas-input basis (CO[P.sub.Gas]) defined as [Heat to GAHP Loop] / [Natural Gas Input]. Combined delivered efficiency, a transient metric summed over a specified time period (e.g. daily), is defined as [Output to space heating (SH) loop + Output to DHW loop] / [Natural Gas & Electricity Input]. Note that when the term "cycle-averaged" is used, this indicates that the total output and total input considered over the entire cycle, including the brief period after the GAHP cycles off and useful heat is extracted from the hydronic loop. Otherwise, when GAHP output and CO[P.sub.Gas] are reported with time, this is a rolling time-averaged value reported at a given moment neglecting this eventual useful heat recovery.
As small errors in loop temperatures can impact estimated GAHP output and efficiencies, the authors performed in-situ calibration of all three pairs of loop temperatures, the supply/return at the GAHP, Space Heating coil, and DHW tank. For example, if the GAHP output was 20,000 Btu/hr with a 1.50 COPGas, a 0.5[degrees]F (0.3[degrees]C) error in supply/return temperatures such that the effect is additive (i.e. difference decreases by 1[degrees]F/0.5[degrees]C), the calculated output would be 15,750 Btu/hr (4.6 kW) and the CO[P.sub.Gas] would be 1.18 for the same firing rate, a 21% error in calculated CO[P.sub.Gas]. Note that the terminology of "supply"/"return" differs for the GAHP and that the GAHP/Space Heat loops are physically connected while the DHW loop is separated by a plate heat exchanger (PHX) (Figure 2). Summarized results are:
* Heating Season #1 - 1st Gen. GAHP: Generally, the GAHP was oversized for the space heating load, which ranged from 0.1-0.3 MMBtu/day (29-88 kWh/day), no more than 3.8 equivalent full-load hours. Over this period, beginning in March, the space heating load was 3X greater than DHW loads, with an average of 52 gal. (197 L) hot water per day consumed. Daily GAHP operation peaks at slightly over 11 hours per day and steadily drops to between 0.5 and 1.5 hours per day as it serves DHW loads only. With these short duration cycles, the GAHP cannot reach a high operating efficiency quickly enough, particularly when the loop temperatures are rising as quickly as they do for a DHW storage tank heat up and where the energy required to heat up the recirculating loop represents a relatively significant load on the GAHP. However, as shown in Figure 3 (Left), the GAHP performed well with combined and space heating (SH) only cycles, often matching performance observed in prior laboratory testing (Glanville, 2017).
* DHW-Only Operation - 1st Gen. GAHP: As per Figure 3 (left figure), the DHW-only CO[P.sub.Gas] values are consistently lower than other operating modes and upon the swap from the 45 gallon Tank #1 to the 80 gallon Tank #2, there is a significant jump in time-averaged CO[P.sub.Ga]s. The larger storage tank, with a larger submerged coil surface area, requires fewer, longer cycles between calls for heat, with less than 12 gallons (45 L) drawn to trigger a reheat for Tank #1 while Tank #2 requires twice that on average. The tank was replaced in conjunction with elimination of the intermediate PHX between GAHP and DHW loops (see Figure 2) and an update of GAHP system controls, specifically sequencing GAHP heating of the main loop then the DHW loop, to limit mixing For all combi systems in general, cycling losses are influenced by the loop fluid volume. For the GAHP, the energy input to bring this loop up to space heating or DHW temperatures is a necessary cost for all on-cycles, particularly for short-duration DHW recovery cycles. For the Tank #2 period, using the calculated energy delivered by the GAHP and that to the storage tank for all cycles, this "load" was estimated as the difference between the two, which is ranges between 4,000 and 6,000 Btus (1.2-1.8 kWh), with a median load of 4,708 Btus (1.4 kWh). For an average temperature rise of the GAHP loop from 75 to 135[degrees]F (23.8 to 57.2[degrees]C), this yields an estimated total loop volume of 9.8 gallons (37.1 L). Figure 4 (Left) shows the estimated delivered efficiency for all on-cycles by prior standby period duration. With loop volume estimate, the improvement in delivered efficiency is shown for a 33% and 50% reduction in the loop volume, using as-measured starting/ending loop temperatures. A significant improvement is shown for these hypothetical reductions in loop volume. The improvement will be lessened for greater DHW loading, resulting in shorter standby periods between cycles. For reference, moving from 1 inch PEX to 3/4 inch would reduce the loop volume by 39.4%.
* Heating Season #2 - 2nd Gen. GAHP: At the start of the 2016-17 heating season, in addition to replacing the GAHP with a 2nd gen. prototype ("Beta 1"), the manufacturer implemented new loop controls--regarding the sequencing of space heating and DHW heating events. The goal was to limit GAHP on-cycles through longer space heating/DHW/combined cycles, as short-cycling heat pumps is detrimental to efficiency, capacity, and reliability. As shown in Figure 3 (Right), a) improved DHW cycling performance was sustained, b) space heating-only cycles showed cycle efficiencies close to but slightly below that of prior lab steady-state testing (e.g. full-fire CO[P.sub.Gas] = 1.44 at 34[degrees]F/1[degrees]C while field unit shows CO[P.sub.Gas,cycle] = 1.0-1.45 for this outdoor condition), c) adjusted controls result in fewer combined SH/DHW cycles overall and drastically reduce DHW-only cycles with outdoor temperatures below 47[degrees]C (8.3[degrees]C). This observed difference between laboratory steady-state and field performance owes to transient effects, system losses (e.g. heat to loop), and shorter cycles. Space heating-only cycles were brief, 48 minutes on average, while combined SH/DHW cycles average 1.7 hours with a peak of 14.7 hours. Concerning comfort, the data suggest DHW was supplied at usable temperatures at all times (>105[degrees]F/40.5[degrees]C) and the delivered air temperature versus outdoor air temperature was steady, using a multi-stage AHU and reset control (Figure 4, Right).
Comparison to Experimental Test Results
As described previously, a 1st gen. GAHP underwent a laboratory investigation, estimating the AFUE and performance at a number of steady state (SS) rating points (Glanville, 2017). In parallel to the aforementioned field investigation, the authors performed extensive steady and dynamic testing of a 2nd gen. GAHP deployed as a simulated combi system, using space heating and DHW loading representative of homes in the Pacific Northwest. Steady rating points use the previous methodology and, for interest of space, the details of the dynamic testing will be reported in subsequent publications. Figure 5 shows an overall comparison of the steady (SS) COPGas, the cycle-averaged COPGas for each cycle during multiple experimentally simulated 24/48-hour loading patterns, and a linear fit of the 2nd generation GAHP field results for combined SH/DHW operation (as per Figure 3 for outdoor temperatures below 47[degrees]C/8.3[degrees]C). Here, good agreement was observed between dynamic experimental and field testing, suggesting a good simulation of dynamic testing, though insufficient detail provided here does not permit a thorough comparison. More salient is the difference between steady and dynamic (experimental and field) operating efficiencies, 0.10-0.15 CO[P.sub.Gas] "points", for reasons previously discussed. Improved system controls and component design may bridge this gap, from steady GAHP efficiency to an as-installed dynamic performance, a topic under current investigation with subsequent laboratory/field evaluation programs. This includes the means of delivering space heating and DHW and also such system features as defrost control, implemented with 2nd gen. units using a "hot gas bypass" methodology. During experimental testing, numerous defrost events triggered "defrost mode", lasting roughly 10 minutes each time. In defrost mode, the system continued heating but at approximately 50% capacity, resulting in a temporary drop in CO[P.sub.Gas] by about 30% while the hot-gas bypass valve flowed warm vapor refrigerant to the evaporator coil for defrost. This loss of capacity is essentially the loss of the "refrigeration effect" while the unit recovers the heat not used for defrost to maintain a fraction of total heating output. Incrementally reducing the impact of these system losses while improving delivered efficiency is critical.
The authors summarized efforts to apply a new residential-scale low-cost GAHP as a combi system. This GAHP combi system delivers up to 80 kBtu/hr (23 kW) for hydronic space heating/DHW with an estimated AFUE of 139%, a 50% improvement over baseline, and an equipment cost target of less than $5,000. By integrating with a forced-air heating system via a hydronic air coil and an indirect storage tank for DHW, the GAHP-driven combi system was shown to meet a home's space and water heating loads simultaneously, through simulated-use testing in a laboratory and through a field demonstration over 15 months at a residence in Tennessee. The authors outlined the performance of the GHP-driven combi system as a function of loading, operating conditions, and system control strategies. Additionally, the impact of system design considerations for component sizing and control were explored. As field trials of these technologies continue towards potential commercialization in the U.S., with more involvement from OEMs and installation contractors, the authors will track new findings against these aggressive equipment cost targets.
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U.S. Census Bureau, 2010. Data referenced from www.census.gov
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|Author:||Glanville, Paul; Suchorabski, Daniel; Keinath, Chris; Garrabrant, Michael|
|Publication:||ASHRAE Conference Papers|
|Date:||Jan 1, 2018|
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