Experimental feasibility study of a new load-based method of testing for light commercial unitary heating, ventilation, and air conditioning (ASHRAE RP-1608)

Abstract

This article investigated the feasibility a new load-based method of test for assessing the comprehensive performance figure of merit of roof top units. The measurements, according to the new load-based approach, characterize the relative changes in the unit performance when it operates in the field as if the unit was installed in an actual building. Depending on the mode of operation and on the outdoor air dry-bulb temperature, the roof top unit efficiency was about 41% to 9% lower than the steady state efficiencies determined by using the ASHRAE 116 Standard if the operation mode of the roof top unit was with economizer closed to the fixed minimum outside air position and without any speed control. When both the economizer and speed control technologies were activated, the measured efficiency increased to 205% with respect to the steady state efficiency. The focus of the load-based test was to provide data that could be used to project energy use and change point linear inverse modeling following ASHRAE Guideline (2014) could be applied to the data to compare expected annual energy use for the baseline (i.e., roof top unit with minimum outside air) to the improved modes of operation with fan speed control and outside air economizer control.

View correction statement:

Corrigendum

Background

In recent years, given the critical importance of energy savings during actual field operation, an integrated energy efficiency ratio (IEER) has been used as a single number figure of merit expressing cooling part-load coefficient of performance (COP) efficiency for unitary air-conditioning and heat pump equipment. Yet, advanced roof top units (RTUs) include a series of accessories, new technologies, and advanced controls that often improve the nonsteady-state energy conversion efficiency. The associated potential energy savings are not properly characterized by steady-state temperature-controlled methods of testing. Simulation and field-testing can provide more information but computational models are difficult to calibrate and field samples are typically small. Field-testing also involves significant delays while waiting for full seasonal weather impacts. Thus, there is a need for developing a standard test procedure in laboratory that can provide a figure of merit of packaged RTUs similar to the miles per gallon figure of merit used in the automotive industry.

When packaged RTUs are integrated in actual buildings, some estimates of the yearly energy performance can be simulated in computer models. However, current models for unitary systems are not well-calibrated to actual system operation. This is mainly due to inaccurate representation of the system behavior at various ambient conditions when it cycled through various states to balance the building indoor thermal loads. The lack of a systematic and comprehensive method of testing for the RTU that includes advanced controls and additional technologies, such as economizer and variable speed fans for example, makes it difficult to develop accurate models for the RTU during actual in-building operation. To address this need, the American Society for Heating, Refrigerating, and Air Conditioning Engineers (ASHRAE) proposed the research project RP-1608 that focused on developing new lab-based performance testing procedures to allow a range of control and other technologies to be tested and system performance determined for multiple climates using regression based hourly projections. The current article summarizes the main highlights of the ASHRAE project RP-1608 and discusses new results of the energy performance of a RTU unit obtained by conducting newly developed load based method of testing in laboratory. A comparison with the energy performance figure of merit from the more conventional steady state and standardized temperature based method of testing in laboratory is also presented in this article.

Literature review

The standards ANSI/ASHRAE 116–2010 (ASHRAE 2010) and ANSI/ASHRAE 37–2009 (ASHRAE 2009a) describe several methods to measure heating and cooling capacities and calculate the system performance using environmental testing rooms. Two of the most important and currently adopted figures of merit are the seasonal energy efficiency ratio (SEER) and the IEER. The SEER depends on outlining system capacity and power profiles over different temperature bins based on tests results. Bin methods are commonly used to estimate building energy consumption (Knebel 1983) and a degradation factor is often included per ASHRAE Standard 116–2010 in order to account for cycling losses during part-load operation (ASHRAE 2010, 2012). The IEER has gained importance in the recent years as a single number figure of merit expressing cooling part-load efficiency. The IEER figure of merit was included in AHRI 340/360-2007 and AHRI 210/240-2008 and it became effective in 2010 as a part of the rating program. The IEER is based on weighted averages of the EER, which is the ratio of the cooling capacity in Btu/h to the power input values in watts at any given set of rating conditions. IEER and EER are both expressed in Btu/W-hours. The IEER provides some improvements with respect to the previously developed figures of merit since units typically operate a significant amount of time under part load conditions (ANSI/AHRI 2007, 2008).

Kavanaugh (2002), expressed concerns that predictions on energy savings based on a single figure of merit could be inaccurate. Comparing AHRI 210/240 tests for two units made by the same manufacturer, Kavanaugh observed that a unit with 18-SEER had only a 6% higher EER at 95°F (35°C) than a 10-SEER rated unit. Furthermore, the author indicated that some factors inflated SEER such as the bin temperatures used for calculation being lower than the actual indoor temperature during 66% of the time. Kavanaugh also mentioned that the external static pressures required in the AHRI Standard are often lower than field-measured data as indicated by Proctor and Parker (2000). Another challenge pointed out by Kavanaugh was that SEER does not have much consideration for dehumidification and how units could be potentially designed with evaporators that are oversized with respect to the compressors used in the systems. These factors reduce power consumption, but can hinder dehumidification capability of the unit due to higher evaporating temperatures. Kavanaugh recommended a new method of testing that does not allow efficiency to be attained by hindering dehumidification capacity, has indoor return temperatures within human comfort zone and uses external static pressure difference values closer to field data. Fairey et al. (2004) mentioned that even though figures of merit might provide a mean for comparison between various equipment, they are usually misunderstood, especially when extrapolating the results across different climate regions. The authors highlighted that there is not any difference for the various climate regions when calculating SEER. The EER at 95°F (35°C) was reported by Fairey et al. (2004) to be of less importance in the SEER calculation when compared to the EER at 82°F (27.8°C), because peak loads are typically found at higher outdoor temperatures. Similar to Kavanaugh (2002), Fairey et al. (2004) also mentioned that the unit static pressure differences are low in the standards method of tests and more realistic values might translate into lower airflow rate through the coil, which, in turn, would increase the energy required in dry climates and decreases performance. With these considerations, Fairey et al. (2004) simulated unit operation across 15 locations in the United States in order to understand the variation of SEER and heating seasonal performance factor (HSPF) with climate conditions and they reported that the SEER could vary by as much as 22% with respect to the reported nameplate value. The hottest climates showed the largest decrease in unit performance with respect to the nameplate SEER.

Field data is of great importance for understanding unitary HVAC actual performance. Jacobs et al. (2003), studied 215 RTUs in 75 newly constructed buildings in California. The goal of the project was to investigate inefficient design and installation of small commercial HVAC systems, which were well-known for consuming more energy than necessary to conditioning indoor environments. In their study, the most commonly found packaged unit size was 5 tons (17 kW), while units between 1 and 10 tons (3.5 to 5 kW) were the most commonly sold (90% of total sales) for new buildings in California. Some challenges were pointed out such as nonworking economizers, improper refrigerant charge, fans that did not provide adequate ventilation air, and simultaneous heating and cooling and thermostats not properly set up. One of the technologies tested in the present article is an outside air (OA) economizer, which was a large fraction of the issues highlighted in Jacobs et al. (2003) study (for example 123 of the 215 units considered in their study had economizers and out of which 64% of the economizers had some type of malfunctioning, including some of them not functioning at all [24%]). Jacobs et al. (2003) reported that 39% of the units airflow rates were measured to be less than 300 cfm/ton as a result of high external static pressure drops, which were reported to be close to 0.5 in. of water. This value of static pressure is much higher than that prescribed by the standard method of tests. Additionally, Jacobs et al. recommended a method of test in laboratory in which economizers are integrated with the RTU being tested in the lab. According to Jacobs et al. (2003) this new test, together with a series of specifications on efficiency, controls, and reliability might attract manufactures to develop high performance small package HVAC units.

Cowan (2004), collected data from four field studies (including Jacobs et al. [2003]), investigating 503 RTUs in 181 commercial buildings, across five states in the United States. Their data were similar to those of Jacobs et al. (2003) and the issues in the units operating in the field occurred with similar frequencies as those originally reported in Jacobs et al. (2003). From this point-of-view, Cowan (2004) confirmed the challenges related to field operation due to refrigerant charge, sensors, thermostats, airflow, and economizers. Forty-two percent of all units had an airflow rate lower than 300 cfm/ton, this was considered too low for airflow rates. Issues with economizers included broken, frozen or missing drive components, OA or mixed air sensor failure, faulty repairs, low changeover temperature set-point, and the use of a single-stage cooling thermostat. Energy savings related with either repairing or replacing economizers, varied from 14% to 40%. Where 14% was associated to adjustments to a working economizer and 40% was associated with repairing or replacing a broken one. An economizer-focused article by Hart et al. (2006), reviewed several field studies, including some of the previously mentioned ones. It their review, they indicated how only about one-third of the potential savings estimated with simulations were being observed in field studies. A report from the Bonneville Power Administration (BPA) describes a packaged RTU servicing pilot and monitoring project conducted in the summer of 2009. About 150 units at 41 sites in the Seattle City Light and Snohomish Public Utility District (PUD) territories were monitored to determine which efficiency and performance measures had the lowest cost and generated the greatest paybacks; which screening methods identified units with the greatest potential for savings; and which methodology to use to create a provisionally deemed energy savings (Hile et al. 2010). Based on the results of the cooling and fan energy savings analysis of the BPA report, a provisionally deemed savings in kWh per ton of cooling was recommended. Hile et al. (2010) also recommended that servicing programs include thermostat measures as these generate greatest savings, and they should consider excluding the system airflow adjustments that are time-consuming to perform and appear to lead to negative savings. The Pacific Northwest Regional Technical Forum undertook an ambitious field study of more than 160 RTUs and developed a protocol (NBI 2012) to project annual energy use based on monitoring. The protocol used an energy signature relative to average daily outside temperature to project energy use in different operating modes. Such a protocol could be adapted for the data collected from the load based lab testing and applied based on a temperature bin methodology.

A more^{Fig. 1, Fig. 2}recent article by Hart et al. (2008) suggested the need to develop an additional testing protocol for RTUs in order to be able to use the results from laboratory tests for models development of RTUs integrated with buildings and in the actual field service operation and for verification of the simulation results. Kavanaugh (2002) and Jacobs et al. (2003) also pointed out interest in developing such new testing protocol for RTUs. Hart et al. (2008) indicated that potential savings outside of the steady state efficiency measurements (SEER, EER, IEER, and HSPF) were not properly captured with current temperature-based methods of test. They emphasized the need for developing a new load-based method of test that could help isolating and quantifying potential energy savings due to advanced controls and additional technologies in the RTUs that act during the cycling and transitory periods when RTUs are operating in the actual buildings. For example, in the Northwest sixth power plan report it was stated that relatively sophisticated HVAC engineering, smart control systems, and careful system operations are needed to reach low-cost HVAC energy savings for the commercial sector. The savings were estimated in the order of 325MW average (NPCC 2010). One of the ways of achieving such savings might be to force higher SEER and HSPF requirements. However, increasing steady state laboratory test based figure of merit efficiencies is not only challenging but also not the dominant factor anymore. Advanced control strategies can provide non-steady energy savings coming from OA economizers, control configuration, fan speed control during ventilation, and cycling. The effect of these strategies on energy consumption cannot be captured by current figures of merit coming from temperature-based laboratory methods of tests and Hart et al. (2008) proposed to look into ways to characterize the RTU performance beyond steady-state efficiencies. To give a broader context to the work described in the present article, an example is taken from the article of Hart et al. (2008). The authors presented results of simulations that evaluated the potential savings of some non-steady state efficiency improvements for RTUs. The study, which covered eight cities in different climate zones in the United States, suggested that there was a potential of 30% to 48% of energy savings from technologies improvements in RTUs. When compared to the 1.5% to 6.7% expected savings from upgrading SEER for example from 13 to 15, it seems that a laboratory method of testing developed ad-hoc to characterize the performance of additional technologies operating as integrated components of an RTU, would be helpful. Hart el al. (2008) also mentioned that since current metrics for energy consumption do not account for the savings potential of advanced technologies, there is little incentive for manufacturers to incorporate them in their units. A test procedure that evaluates total system performance and provides means for reliably comparing these technologies would allow equipment with better performance and energy savings to be well-recognized and promoted. Their article recommended a comprehensive laboratory test that included ventilation and other control improvements. Their proposed idea for a new method of test was that it should account for economizer effectiveness, ventilation damper losses, control issues, cycling losses, ventilation operation during warm-up, the use of condenser pre-cooling, or the impact of resistance heat on heat pump effectiveness. ASHRAE, recognizing the broad impact that such a load-based method of test could have in the HVAC community, proposed a research project RP-1608 that focused on developing a new lab-based performance testing procedures to allow a range of control and other technologies to be tested and system performance determined for multiple climates using regression based hourly projections. The main goal of the work presented in this article was to investigate the feasibility of a new load-based method of test for assessing the comprehensive performance figure of merit of RTUs for light-commercial building applications. Six outdoor dry-bulb temperatures ranging from 57°F (13.9°C) to 105°F (40.6°C) were selected for testing the RTU at different outdoor conditions that would yield six different indoor cooling loads. The dry-bulb temperatures and their corresponding wet bulb temperatures taken for the same hours were obtained from TMY3 climate zone data for Salem, OR, which is representative for climate zone 4C (that is, marine climate zone) and for Memphis, TN, which is representative for climate zone 3A (that is, moist climate zone). The unit static pressure was set to 1.0 in wg. (∼249 Pa), which was higher than the settings called for by ANSI/AHRI Standard 340/360. The reason for this deviation was that field studies such as Jacobs et al. (2003) reported that average static pressure drop across typical RTUs is higher than 0.3 in wg.

Experimental methodology and test facility

The experiments carried out for investigating the feasibility of the new method of test developed in the present report were conducted in a large-scale psychrometric chamber (Cremaschi and Lee 2008), which consisted of two adjacent rooms and each room had dimensions of 19 ft (5.8 m) length by 21 ft (6.4 m) width and 16 ft (4.9 m) ceiling-to-floor height and capacity up to 20 tons (70 kW) of refrigeration. The equipment used for the experiments was an electrically driven RTU with 15-ton (53-kW) nominal capacity. Figures 1 and 2 show the schematic of the RTU installed inside the psychrometric chamber. The RTU had two digital scroll compressors, a variable speed supply air blower, and associated variable speed drives. For the present work, an economizer was installed on the left side of the RTU unit and supply, return, mixed and OA temperatures were monitored with the sensors placed in the locations shown in Figure 3. The ASHRAE RP-1608 Final report (Cremaschi and Perez 2016) provides the details of the facility, specific procedures for setting the loads, thermostat setting sequences, and defining the data analysis and reduction. Table 1 provides a summary of the RTU specifications.

Fig. 1. Schematic floor layout of the RTU under testing when it was installed inside the psychrometric chamber facility; locations of the main measuring sensing points, and definition of the conditioning bays of the chamber.

Fig. 2. Photos of the 15-ton (53 kW) nominal capacity air-to-air RTU used and installed inside the psychrometric chamber.

Fig. 3. Side view schematic the RTU with economizer when it was installed in the OSU outdoor room chamber for the tests of the present study (RA: return air; SA: supply air; OA: outside air; and MA: mixed air).

Table 1. Specifications of the roof top unit (RTU) used.

	Roof top unit/air-to-air heat pump
Unit type	electrically driven
Nominal capacity	15-tons (53 kW) of refrigeration
Refrigerant	R-410A
Nominal airflow rate	3600 cfm
Supply air fan	460 VAC – 3 phase – 5hp (3.73 kW; variable speed)
Condenser fans	2 – 460 VAC – 3-phase – 560 W
Compressor	2× Digital scroll compressor. Nominal power: 6 hp
Expansion valve	1 thermal expansion valve per stage (2 for each stage of heating and 2 for each stage of cooling)
Economizer	Gear driven, sensible controlled, opposing blades, with manometric exhaust

The type of control sequence used can be referred to as single sensor dry-bulb temperature changeover. The high limit lockout sequence was the one that determined when the OA was too hot for the economizer to open. The mixed air temperature was also monitored. Very low mixed air temperature results in extremely low supply air temperatures, which can cause occupants to feel discomfort (Hart et al. 2006). When this was the case, the economizer returned to the minimum OA position.

Test conditions and experimental methodology

Six outdoor dry-bulb temperatures ranging from 57°F (13.9°C) to 105°F (40.6°C) were selected for testing the RTU at different outdoor conditions that would yield six different indoor cooling loads. The dry-bulb temperatures and their corresponding wet-bulb temperatures taken for the same hours were obtained from TMY3 climate zone data for Salem, OR, which is representative for climate zone 4C (that is, marine climate zone) and for Memphis, TN, which is representative for climate zone 3A (that is, moist climate zone). The unit static pressure was set to 1.0 in wg. (∼249 Pa), which was higher than the setting called for in the Standard ANSI/AHRI 340/360.

Eighty-four tests were conducted to develop a new load based method of test that captured well the relative changes in capacity or in efficiency due to unit modifications, for example, airflow, economizers, etc. However, it should be emphasized that measuring a specific capacity and efficiency in one laboratory and attempting to replicate similar results at another laboratory test facility for each test point might be unfeasible. Verifying lab to lab repeatability was not part of the current research.

Time-based and cycle-based averaging approaches were presented as good ways to conduct the data analysis and data reduction and for details the authors reference the ASHRAE RP-1608 final report (Cremaschi and Perez 2016). Figure 4 shows an example of a complete load load-based cooling test from start period of the test, in which the RTU run steadily in the chamber and both outdoor and indoor rooms were operated under temperature controlled modes, i.e., according to the steady state ASHRAE 116 Standard method of test. A cycle-based approach with a minimum of two cycles to be recorded was recommended as the best procedure to be used in a load-based method of testing. The thermal inertia of the particular facility and of this particular unit was the basis to selecting the duration of the start-up time, the duration of the transition from temperature-based to load-based operation, and the minimum recording time of the test. The start-up time was 30 min, the transition time from temperature-based to load-based operation of the facility housing the RTU was 60 min, and the minimum recording time was two cycles after the first cycle of the RTU, which was approximately 45 min. Table 2 shows all the measured and calculated variables required for each load-based test. Table 2 includes all the data required by ASHRAE 116–2010 Standard plus a series of parameters inherent to the newly developed load-based method of test and the technologies that were tested.

Fig. 4. Example of a complete load-based cooling test from startup of the RTU (both outdoor and indoor rooms were temperature controlled during the initial steady state part of the test, then the indoor room was switched to load controlled mode at the beginning of the transition period; the actual recording period started at the end of the transition period, referred as to “load-based test recording” section).

Table 2. Variables recorded and calculated.

Item	Units
Date
Barometric pressure	Inches of water (Pa)
Times
Power input to equipment(a)	W
Frequency	Hz
External resistance to airflow or coil pressure drop	Inches of water (Pa)
Fan speed(s)	Rpm
Dry-bulb temperature of indoor return air entering equipment	°F (°C)
Wet-bulb temperature of indoor return air entering equipment	°F (°C)
Dry-bulb temperature of supply air leaving equipment	°F (°C)
Wet-bulb temperature of the supply air leaving equipment	°F (°C)
Dry-bulb temperature of outdoor return air entering the economizer	°F (°C)
Wet-bulb temperature of outdoor return air entering the economizer	°F (°C)
Throat diameter of nozzle(s)	In (cm)
Static pressure difference across nozzle (s)	in. water (Pa)
Temperature at nozzle throat	°F (°C)
Pressure at nozzle throat	in. water (Pa)
Static pressure difference across economizer inlet	in. water (Pa)
Power to electric heaters	kW
Steam output from the humidifier	lb/h (kg/s)
Cycling time (if any cycling is present)	min
Supply airflow rate	CFM (m^3/s)
Outside airflow rate through economizer	CFM (m^3/s)
Total capacity of the unit	BTU/h (kW)
Sensible capacity of the unit	BTU/h (kW)
Total capacity of the unit without outdoor air ventilation effects
Total coefficient of performance
Sensible coefficient of performance
Total coefficient of performance without outdoor air ventilation effects
Temperature control figure of merit	°F (°C)

The effects on the RTU overall figure of merit from the economizer and control fan speed technologies were assessed during the tests as these modes of operation were individually activated and ran with the RTU in the laboratory setup. The modulating capacity capability of digital scroll compressor technology installed in the RTU was also tested in addition to the economizer and variable speed fan technologies. The influence of such modulating capability on unit efficiency in terms of energy consumption and return temperature control was investigated. Three methods were used to cycle and to control the unit under testing and they were defined as follows:

•	Method M1: The compressor was cycled on and off, and the economizer was opened or closed according to the thermostat settings; however, the fan speed remained constant at its low value of 66%. The reason not to modify the fan speed was to introduce one variable change at a time while progressively developing the economizer control procedure explained in the thermostat settings for cooling tests.
•	Method M2: The compressor operated based on its fully modulating capacity and the economizer was opened or closed according to the thermostat settings, while the fan speed remained constant at 66% for the same reasons as in method M1.
•	Method M3: The compressor was cycled on and off, the economizer was opened or closed and the supply air fan speed was adjusted according to the thermostat settings. Method M3 introduced all the variables and set-points for the economizer and speed control fan and the unit run in the chamber as it would operate in the actual field.

The instantaneous cooling capacity and power consumption were measured at a sample rate of 2 s. The thermodynamic properties equations were taken from Hyland and Wexter (1983a, 1983b). The average COPs were calculated as the ratio of these two instantaneous quantities. In addition, the COPs were integrated in time and averaged with respect to the time of the recording period. This was referred as integrated average COP or simply COP in the present article and it was the figure of merit that represented the unit energy efficiency when the unit operated under load controlled testing conditions.

Uncertainty analysis

Based on^{Fig. 5}the technical approach and instrumentation accuracy, an uncertainty analysis was conducted according to the uncertainty propagation method suggested by Taylor and Kuyatt (1994) and the results are summarized in Tables 3 and 4. The uncertainty, repeatability, and degree of confidence of the load-based tests results were quantified by conducting five uncertainty estimation tests, seven repeatability tests, and two load perturbation tests. The uncertainty was estimated at 3.5% for the total load and 4% for the COP. These values of uncertainty were also confirmed with repetition tests, which yielded total loads variation within 3.5% (or less) and COPs variation within 4% to 5.2%. Results from the variable tests suggested that when ±15% of load fluctuations were present during a test the final average calculated COPs were still representative of the COP of the unit under constant load. The error in estimating the COPs from averages during tests with some load perturbations by as a much as ±15% were up to ±4% with respect to COP with constant loads.

Table 3. Sensitivity analysis towards uncertainty propagation for load and COP.

Measured variable	Sensible load	Total load	COP sensible	COP total
Static pressure across flow nozzles	5.8%	1.3%	5.7%	1.3%
Indoor dry bulb	46.6%	0.0%	45.7%	0.0%
Indoor wet bulb	0.5%	54.6%	0.5%	54.4%
Supply air dry bulb	46.8%	0.0%	45.8%	0.0%
Supply air wet bulb	0.3%	44.1%	0.3%	43.9%
Unit power	0.0%	0.0%	2.0%	0.5%

Table 4. Results from the experimental uncertainty analysis of the experiments.

Calculated variable	Uncertainty
Sensible load	±2.91%
COP sensible	±1.98%
Total load	±3.50%
COP	±4.04%

The psychrometric chamber absolute measurements for airflow rates and heat transfer capacity were verified by Worthington (2011) and they were within 5% of the capacity and 2% of the airflow rate obtained at the outside laboratories. Corti and Marelli (2013) also tested the 15-tons (53 kW) RTU of the present study at several indoor and outdoor temperature conditions and they reported a heat balance between the air and refrigerant sides that ranged from −18.9% to 7.7%. The discrepancy between airside and refrigerant side was due to the fact the indoor coil outlet refrigerant temperature (i.e., the evaporator outlet refrigerant temperature when the RTU run in cooling mode) was assumed to be equal to the compressor suction line temperature. Only the latter was directly measured with an in-line t-type thermocouple in their work.

Discussion of the experimental results

The new testing approach developed aims to characterize the unit performance at specified outdoor dry- and wet-bulb temperatures and specified building space sensible and latent loads. The specified sensible and latent loads were obtained from simulations and they were based on ASHRAE Standard 90.1 of typical building requirements. Pacific Northwest National Laboratory (PNNL) provided the prototype simulations for a small office building of approximately 5500 ft² (511 m²) floor area and these simulations were used for estimating the cooling loads to be set in the psychrometric chamber. The climate zones used for the simulations were zone 4C–Salem, OR and zone 3A–Memphis, TN. These two locations were selected because the resulting loads from the simulations allowed testing the RTU under part-load conditions for a broad range of outdoor temperatures. Figure 5 provides the sensible cooling loads measured and Figure 6 shows the associated COPs. The test conditions and legend of each experiment are summarized in the table below Figure 6.

Fig. 5. Experimental results of the sensible load versus outdoor dry-bulb temperature in cooling mode of the RTU.

Fig. 6. Normalized sensible COP versus unit operating modes at outdoor temperature of 65°F (18.3°C; that shows that the newly developed load-based MOT captured the relative changes of efficiency of this unit operating mode with respect to steady state efficiencies).

Two main groups were observed in the Figure 5: The ones with regular load, which follow the red solid trend-line, and the ones with reduced load following the blue dashed trend-line. The reduced loads were about half of the regular loads and they focused on testing the RTU when it was subjected to severely reduced part load conditions. Sensible loads increased as the outdoor air temperatures increased. The measured sensible loads were within ±5118 BTU/h (±1.5 kW) of each other for a particular temperature and load condition. The reason for this tolerance, shown for two representative data points in Figure 5, is that the sensible load provided by the electric heaters increased discretely in factors of 6824 BTU/h (2 kW). That is, on top of the fan power, the sensible load in the indoor room could be increased by 2, 4, and 6 KW, etc., by using the heaters. Tests outside the ±5118 BTU/h (±1.5 kW) tolerance were considered outliers and are indicated as such in the results. The error due to the repetition a particular test with the exact same load, temperature and technology used was smaller than the tolerance of the experiments.

The latent load was selected for the same office space. PNNL prototype simulations were available for the office space and the latent load was calculated based on the number of people present in the office. The data showed that 31 people occupied the office during 0800 to 1700. The ASHRAE Handbook of Fundamentals (2009b) recommends a latent load per person of 250 Btu/h for an office with moderately active office work (when the indoor design temperature is 80°F (26.7°C). Thus, the total latent load was estimated according to Equation 1(1) and was constant throughout the tests of the present work. (1)

A different airflow rate than at full fan speed tests changed the amount of humidity being introduced to the indoor room and a variation of the amount of air circulating back to the evaporator resulted in a variation of the amount of unit dehumidification. The combination of these two effects, modified the humidity balance between indoor and outdoor rooms, therefore, the latent load did not necessarily have be the same as in full fan speed tests. In other words, a speed control test or an economizer with speed control test did not necessarily have the same total cooling loads as the corresponding full fan speed test. To sum up, for every test, the sensible load was set as close as possible to the desired condition, regardless of the technology used. Whereas the latent load was a combination of the constant space load set by the steam wand and the humidity that entered the room through the economizer damper, and is, therefore, inherent to the technology being tested.

The effects on the RTU overall figure of merit from the economizer and control fan speed technologies were assessed during the tests as these modes of operation were individually activated and run with the RTU in the laboratory setup. As an example, Figure 6 shows the normalized sensible COP of the unit operating in different modes for one outdoor temperature of 65°F (18.3°C). For the normalization of the efficiency results, the sensible COP of the RTU when tested according to the conventional ASHRAE 116 standard (i.e., AHRI 340/360) in cooling mode and at 82°F (∼27.8°C) outdoor air temperature, was selected to be the baseline figure of merit for efficiency of the unit. Then, the normalized sensible COPs in Figure 6 were calculated as follows: (2)

Depending on^{Fig. 7, Fig. 8}the mode of operation, the resulting measured performance of the RTU varied from normalized sensible COP of 67% (i.e., about 33% efficiency decrease with respect to ASHRAE 116 steady state efficiency tests) for mode 1 and up to COP of 297% for mode 10, i.e., about three times higher efficiency than that reported for steady state efficiency tests.

A cycle-based approach with a minimum of two cycles to be recorded was recommended as the best procedure to be used in a load-based method of testing. Although it was found that for long periods if more than two cycles had occurred, time based approach was also valid. Results are summarized in Figure 7, which shows the normalized COP as a function of time for a series of tests conducted at 72°F (22.2°C) and full speed fan.

Fig. 7. Minimum test recording period for a load-based MOT (with an example of a normalized COP versus time—series B: 72°F [22.2°C] full fan speed test).

The horizontal blue dashes show the cycling temperature profile for the test. From left to right, temperature decreasing indicates mechanical cooling is on, while temperature increasing indicates mechanical cooling is off. The cycling of cooling stages was done according to the thermostat settings. Five different starting points for COP calculation are presented in Figure 7 and, in order to compare the five ways of calculating COP, their COP were normalized with respect to the total COP calculated from the beginning to the end of the 90 min test. The average cycle duration was τ = 21 min and 4.3 cycles took place. Figure 7 shows that regardless of the starting point, all averaging approaches show large peaks or valleys within the first and second cycles of recording for the calculation of the average COP. After a period 2τ (of 42 min for the test in Figure 7), the integrated COP/COP_total tends to 1, which means that the measured integrated COP calculated during the period of 2τ is a good representation of the efficiency performance figure of merit of the unit for the load-based method of test.

Figure 8 shows experimental results of the normalized total measured COP of the unit by using the newly developed load-based method of test and for three modes of operation: series B: full fan speed with minimum OA; series C: fan speed control; and series D: economizer with speed control tests. For the normalization of the efficiency results, the baseline total COP of the RTU was selected when the unit was tested according to the conventional ASHRAE 116 Standard (i.e., AHRI 340/360) in cooling mode and at 82°F (∼27.8°C) outdoor air temperature. Depending on the mode of operation and the outdoor air dry-bulb temperature, the resulting measured total COP varied from 59% to 205%. In Figure 8, this means that the RTU efficiency was about 41% to 9% lower than the steady state efficiencies determined by using the ASHRAE 116 Standard if the operation mode of the RTU was with economizer closed to the fixed minimum OA position and without any speed control. When both the economizer and speed control technologies were activated in this particular RTU, the measured efficiency increased to up 205% with respect to the steady state efficiency coming from the test conducted at ASHRAE 116 Standard conditions. While an energy estimating method was not part of the current study scope, Figure 8 illustrates that the load-based test was able to provide data that could be used to project energy use. Change-point linear inverse modeling following ASHRAE Guideline (2014) could be applied to the data to compare expected annual energy use for the baseline (i.e., mode B: RTU with minimum OA) to the improved modes of operation with fan speed control (mode C) and OA economizer control (mode D).

Fig. 8. Trends of the normalized total COP versus outdoor dry-bulb temperature for three different unit operating modes (series B, C, and D with RTU run in cooling mode). This plot shows with an example how the relative changes in the efficiency were captured by the load-based method of test and how these results can be used.

Conclusions

This article investigated the feasibility a new load-based method of test for assessing the comprehensive performance figure of merit of RTUs for light-commercial building applications. The newly developed load-based method of test was suitable to implement in existing psychrometric chamber facilities with minor modifications of the instrumentation typically used in those facilities. It was important that the test conditions allowed the unit to be tested under part-load operation and for a broad range of outdoor temperatures. Six outdoor dry-bulb temperatures ranging from 57°F (13.9°C) to 105°F (40.6°C) were selected and the unit static pressure was set to 1.0 in wg. (∼249 Pa), which was higher than the ones set in the Standard ANSI/AHRI 340/360.

A cycle-based approach with a minimum of two cycles to be recorded was recommended as the best procedure to be used in a load-based method of testing. Depending on the mode of operation and on the outdoor air dry-bulb temperature, the RTU efficiency was about 41% to 9% lower than the steady state efficiencies determined by using the ASHRAE 116 Standard if the operation mode of the RTU was with economizer closed to the fixed minimum OA position and without any speed control. When both the economizer and speed control technologies were activated in the RTU, the measured efficiency increased to up 205% with respect to the steady state efficiency coming from the test conducted at ASHRAE 116 standard conditions.

Nomenclature

COP	=	unit total coefficient of performance, dimensionless
COPsens	=	unit sensible coefficient of performance, dimensionless
MOA	=	minimum outside air (−)
MOT	=	method of test (−)
OA	=	outside air (−)
	=	latent load set by the testing facility in the indoor room, BTU/h (kW)

Acknowledgments

The authors acknowledge and thank ASHRAE Technical Committee TC8.11 Unitary and Room Air Conditioners and Heat Pumps and the Project Monitoring Committee Members Reid Hart (chair, PNNL), Michael Deru (NREL), Don Schuster (IRCO), Jay Baggett (Trane), Chris Stone (AHRI), and Kristin Heinemeier (UCDAVIS) for providing technical consultation on this project. The authors would like to thank AAON Inc., for providing the equipment and components used in the experimental campaign of the present work.

Funding

The authors acknowledge and thank the American Society of Heating, Refrigerating, and Air Conditioning Engineers (ASHRAE) for supporting this work.