© 2004. American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). Published in ASHRAE Transactions, Vol. 110, Part 1. For personal use only. Additional reproduction, distribution, or transmission in either print or digital form is not permitted without ASHRAE’s prior written permission.

AN-04-5-3

Heat Pump System Performance in Northern Climates Paul W. Francisco

David Baylon

Member ASHRAE

Bob Davis

ABSTRACT Single-value ratings for heat pumps and air conditioners have become industry standards and are widely used by consumers, designers, and program managers for equipment selection and prediction of performance. On the heating side, the standard rating is the heating seasonal performance factor (HSPF), and on the cooling side the standard rating is the seasonal equipment efficiency ratio (SEER). These values are determined under set conditions. However, the actual performance of the equipment depends on the climate in which it is being used. Furthermore, the overall energy use can be greatly impacted by factors such as control strategy and duct losses. This paper presents the results of computer modeling using the bin method to estimate the impact of climate, certain common control strategies, sizing approaches, and duct losses on the HSPF of heat pumps in two climates in the northwest United States. Comparisons with field data and observations show impacts on heat pump performance consistent with modeling results. INTRODUCTION As energy prices increase, people are paying more attention to the seasonal efficiency of heat pumps and air conditioners. This efficiency is often characterized by a single number, which is determined at a specific set of test conditions. For heat pumps, this number is the HSPF (heating season performance factor), and for air conditioners the number is the SEER (seasonal energy efficiency ratio). These values are used by consumers to compare conditioning systems, by program managers and regulators to predict savings and establish incentives, and by designers to specify equipment.

Larry Palmiter

The SEER rating uses the results of operating the air conditioner at three different conditions (ARI 1994). All of them are done at 82°F outdoor temperature and 80°F indoor temperature at the indoor coil, for a sensible load temperature difference of only 2°F. Two of the tests are done with low indoor humidity (dry coil)—one of these is done at steadystate conditions (typically about 10 to 15 minutes after the unit has begun operation); the other is done with the compressor operating for 6 minutes and then off for 24 minutes. These two results are used to determine the part-load factor, which accounts for losses due to the cycling of the compressor. The third test is done with higher indoor humidity such that the coil is wet and is also done under steady-state conditions. The SEER rating is the EER (energy efficiency ratio, defined as the ratio of the compressor output in kBtuh to the power input in kW) for this third test multiplied by the part-load factor. The fact that the SEER rating is higher than the commonly published EER rating is because the commonly published EER rating test is performed at 95°F outdoor dry-bulb. The HSPF uses a similar technique to get the part-load factor as that used in determining the SEER (ARI 1994). The HSPF test, however, uses a much broader range of conditions to get the rating. Compressor efficiencies at both 17°F and 47°F are used, as well as the defrost penalty at 35°F, using a 90-minute defrost cycle. The data are applied to multiple design loads and in multiple climates, using a bin calculation technique. This results in many different HSPF ratings for each piece of equipment. The rating that is published is based on the U.S. Department of Energy Climate Region IV using the minimum design load. Despite the broad use of these single-value ratings, the actual seasonal performance of a specific piece of equipment

P.W. Francisco is a research specialist at the Building Research Council, School of Architecture, University of Illinois, Champaign. D. Baylon is president, B. Davis is a research scientist, and L. Palmiter is senior scientist at Ecotope, Inc., Seattle, Wash.

442

©2004 ASHRAE.

© 2004. American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). Published in ASHRAE Transactions, Vol. 110, Part 1. For personal use only. Additional reproduction, distribution, or transmission in either print or digital form is not permitted without ASHRAE’s prior written permission.

can be very different from the ratings provided. One of the primary causes of this is that the conditions under which the rating tests are performed may not represent the location of interest. For example, if the equipment is located in a climate that differs significantly from the climate selected for publication of HSPF values, the seasonal efficiency of the heat pump may be very different from the stated rating. The differences between a specific climate and the rating conditions can result in either an improvement or a reduction in seasonal performance. This fact is well known among experts, but the way in which these ratings are used suggests that this information is not common knowledge among users, practitioners, and policy-makers (for example see EPA 1998 and ICC 1998). Another factor that can affect the actual performance of a heat pump is the control strategy, especially on the heating side. The point at which the backup heating comes on can have a major impact on the actual energy use of a heat pump, since the efficiencies of electric resistance (100%) and natural gas combustion (typically about 80% for standard gas furnaces, 95% for condensing types) are both much less than the efficiency of a compressor (capacity divided by power input, usually 200% to 400% at the standard rating points). Averaging the backup contribution into the overall performance will reduce the apparent equipment efficiency. Backup heat is considered in HSPF ratings, with the assumption that the backup heat only comes on when necessary. Experience with the manner in which systems get installed in the Pacific Northwest shows that the desire to ensure customer comfort or prevent customer complaints results in significantly greater use of backup heat. Duct losses can also have a major effect on the system performance. Duct losses cause the system to run longer than it would otherwise and on the heating side can cause the backup to be required at warmer temperatures than if there were no duct losses. This effect was evaluated for air conditioners under hot conditions by Walker et al. (1998), who found that, even with good sizing and installation, the delivered cooling under the conditions evaluated could be as much as 50% lower than the rated output. Palmiter and Francisco (1997) showed that duct losses in the heating mode had a larger impact on the system efficiency for heat pumps than for furnaces, with the efficiency loss often double that for furnaces for the same duct leakage. This is because delivery temperatures are much lower for compressors than for furnaces and extra load is often made up by backup heating elements. This paper describes the results of computer simulations to investigate the performance of heat pump systems for given ratings, installation practices, and duct systems in two locations in the Pacific Northwest—Seattle and Spokane, Washington. These computer simulations use a bin method calculation, similar to that used in the determination of the HSPF ratings, to estimate the efficiency of the heating system at each bin and then combine the results to provide a seasonal efficiency calculation. No cooling season results are reported in this paper, though a similar method can be applied. Results ASHRAE Transactions: Symposia

of both heating and cooling simulations are included in a previous report (Francisco et al. 2002). The results are expressed as overall seasonal efficiencies. There are three pertinent efficiencies. The first is duct efficiency (the ratio of the heat output required with no ducts to the heat output with the ducts). The second is the equipment efficiency (ratio of the heat energy output to the operating energy input). The equipment efficiency is also affected by the duct losses because the duct losses change the amount of time that the equipment runs in each bin. The last efficiency is the heating seasonal system efficiency (referred to as the system efficiency), which is the product of the duct efficiency and the equipment efficiency. System efficiency is also the ratio of the house heating load (seasonal) to the actual operating energy input, i.e., the amount of energy that was actually required to keep the house warm throughout the year. It is this last efficiency that actually describes how well the system as a whole is performing and is an index for how much it will cost to heat the house as well as a baseline for potential savings calculations. Note that the duct efficiency as defined here is the same as the distribution efficiency in ASHRAE Standard 152 (ASHRAE 2003), with the exception that in Standard 152 the distribution efficiency includes an “equipment efficiency factor” that here is included in the equipment efficiency. SIMULATION PARAMETERS The following sections describe the model and the combination of factors affecting heat pump performance that are considered in this paper. These include climate descriptions, building descriptions, duct loss characteristics, heat pump characteristics, and different control strategy options. Simulation Model The analysis in this paper is based on computer simulations. These computer simulations use a bin method calculation to calculate the system efficiency at each bin. The model includes a specification of the building load at each temperature bin, explicit performance characteristics from the manufacturer’s literature, the effect of any backup resistance heat required to meet the building load at each bin, and the impact of duct losses. The duct calculation includes the regain of useful heat that is initially lost to unconditioned spaces but recovered to the heated zones and the interaction of unbalanced leakage with natural infiltration. In each bin, the compressor supplies as much heat as possible under the specified control strategy, with enough resistance heat added to supply any shortfall. The results from each bin are multiplied by the number of hours in the bin and then summed to get annual heating season results. These annual results are used to get the efficiencies described previously. The bin method is a generally recommended method for evaluating seasonal energy performance. This is the method recommended in ACCA Manual H (ACCA 1984) and described in McQuiston and Parker (1988). The method is also recognized by ASHRAE (Knebel 1984, which also outlines 443

© 2004. American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). Published in ASHRAE Transactions, Vol. 110, Part 1. For personal use only. Additional reproduction, distribution, or transmission in either print or digital form is not permitted without ASHRAE’s prior written permission.

Figure 1 Comparison of Spokane climate to other locations.

the general technique). The duct model adapts the ASHRAE Standard 152 equations to the bin model format to calculate overall distribution efficiency. Climate Descriptions Seattle has a fairly mild climate, with a small number of hours below freezing. The average temperature during the months of October to April is about 46°F. Due to high humidity levels during the heating season, there is often a significant need for defrost. On the other hand, there are relatively few hours in temperature bins below 32°F. In most ways the Seattle climate is favorable for the use of heat pumps. Spokane, which is located in eastern Washington, has a large number of hours below freezing, where heat pump performance drops off considerably and the need for backup heating increases. The average temperature during the months of October to April is about 36°F. As seen in Figure 1, the heating season in Spokane is comparable to eastern locations such as Boston and Philadelphia (which is typical of the climate used for the standard HSPF rating) and Rocky Mountain locations such as Salt Lake City. This graph shows the hours in each bin below the balance point of the prototype houses (55°F) as a fraction of the total number of hours below the balance point. This method of describing the climate is the way in which climates are assigned climatic regions for purposes of determining HSPF. To reinforce the extent to which Spokane looks like these locations, Minneapolis is also shown. Seattle is also shown to emphasize how favorable Seattle is for heat pumps. Building Descriptions Two prototype buildings were evaluated. The first is a 1350 ft2 single-story house built over a vented crawlspace. The floor and the wall are insulated to R-19, the attic to R-38. The overall house UA is 377 Btuh/°F. In the model, the crawlspace is thermally connected to the ground temperature, which is set 444

to the annual average temperature. The return ducts are all in the attic, and the supply ducts are all in the crawlspace. Both sides were assumed to have 10% duct leakage to outside as a percentage of air-handler flow. The supply ducts were assumed to have R-4 insulation, with the return ducts uninsulated. This is common practice in the Pacific Northwest and would also apply to many cases where a heat pump is being added to an existing building. There are codes that require greater insulation than that used in the modeling; in these cases, the duct losses will be lower if the insulation is installed well. The supply duct surface area was 365 ft2. Because of the single return system, the return ducts were assumed to have a surface area of about 80 ft2. For a 3.5 ton heat pump with a system flow rate of 1400 cfm, this results in a return duct conduction efficiency of 96.5% if the ducts are uninsulated, using the equations in ASHRAE Standard 152P (ASHRAE 2003). Reducing the airflow rate to 1000 cfm would provide a conduction efficiency of 95.2%, and increasing the surface area to 100 ft2 (with 1400 cfm) gives a conduction efficiency of 95.7%. Because the return ducts have a conduction loss of less than 5%, so the effect of insulating these ducts on the overall efficiency is small. The second building is a 2184 ft2 split-level house, with one-half of the house being built over a conditioned basement and the other over a crawlspace. The house UA is 565 Btuh/°F. The return ducts are entirely within the conditioned space, and the supply ducts are half in the attic and half between the floors. The supply ducts are assumed to have 5% duct leakage to outside and are insulated to R-4, and the return ducts are assumed to have no leakage or conduction losses. The supply duct surface area outside the conditioned space is 205 ft2. Heat Pump Characteristics and Control Strategy Options Two 3.5 ton heat pumps were considered. The first has an 8.2 HSPF rating, while the second has a 7.2 HSPF rating. Performance characteristics were taken from manufacturer’s data and include the defrost penalty. The system airflow rate for both units is 1400 cfm, which is the standard recommendation of 400 cfm/ton. The 3.5 ton heat pump is oversized for the buildings described. This is an artifact from a preliminary study that led to the project from which these results originated (Francisco et al. 2002). In that project, a variety of building insulation levels were considered, and the heat pump was sized for the building with the largest load. In order to make comparisons across building insulation levels simpler, the same heat pump was used for all cases. An evaluation of the effect of sizing showed that changing the unit to 2.5 tons had about a 5% effect on the overall seasonal system efficiency and that the control strategy and duct losses determined whether the change was an increase or a decrease. This is discussed further in the “Results” section. In all cases, a degradation coefficient of 0.25 was used to account for compressor cycling losses. This is the U.S. DepartASHRAE Transactions: Symposia

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Table 1. Bin Temp. (°F)

Compressor Capacity and Resistance Use for Different Control Strategies

Best Case 1

Comfort Assist 2

1

Low Ambient Cutout

2

Comp. (Btuh) Res. (kW) Comp. (Btuh) Res. (kW)

1

Comp. (Btuh)

2

Plus 5 1

Res. (kW) Comp. (Btuh) Res.2 (kW)

52

44050

0

44050

0

44050

0

44050

0.15

47

40700

0

40700

0

40700

0

40700

0.43

42

36150

0

36150

0

36150

0

36150

0.76

37

31600

0

31600

0

31600

0

31600

1.17

32

28850

0

28850

0

28850

0

28850

1.61

27

26100

0

26100

0

0

4.01

26100

2.12

22

24450

0

24450

2.64

0

4.79

24450

2.64

17

22800

1.43

22800

3.23

0

5.60

22800

3.23

12

21050

2.91

21050

3.89

0

6.42

21050

3.89

7

19300

4.40

19300

4.66

0

7.27

19300

4.66

2

17600

5.86

17600

5.86

0

8.15

17600

5.86

-3

15900

7.33

15900

7.33

0

9.05

15900

7.33

-8

14200

8.80

14200

8.80

0

9.97

14200

8.80

-13

12500

10.27

12500

10.27

0

10.93

12500

10.27

-18

10800

11.74

10800

11.74

0

11.91

10800

11.74

-23

9100

13.21

9100

13.21

0

12.93

9100

13.21

1. Compressor capacity available for heating whenever heat pump is on. 2. Average kWh used per hour of runtime.

ment of Energy (DOE) default value. Kao et al. (1987) performed a review of the suitability of this assumption. Following a review of this report, the default value was considered acceptable for this work. Other values suggested as possibilities by Kao et al. (1987), including 0.30 and 0.34, would have small impacts on the order of 3% to 4% in those bins when the compressor is cycling, with the result being to increase the energy required by the heat pump to meet the setpoint. Later research has indicated that a smaller degradation coefficient may be appropriate for cooling loads (Henderson et al. 2000), which may also apply to heating loads and would reduce the losses due to cycling. Manufacturers may use measured values for degradation coefficients, but no different number was available for this study, so the default was used. No cycling losses were accounted for with regard to the ducts; the calculations were done assuming steady-state duct operation. This results in a slight overestimation of efficiency. According to ASHRAE Standard 152 (ASHRAE 2003) cycling losses are about 2% for flex ducts and 5% for metal ducts. The use of setback thermostat control is also not considered. If the thermostat setpoint is set back for part of the day, and the thermostat is not designed to recover from setback using the compressor, the result will be an increase in resistance and a reduction in the overall system efficiency. ASHRAE Transactions: Symposia

Four different control strategy options were considered for each case. Table 1 shows a comparative example of these strategies. The first is the “best case,” with the compressor providing all of the heating possible and backup resistance filling in as necessary to meet the load. In this situation, the compressor runs constantly in those bins that are cold enough that backup is needed. It is called the “best case” because this provides the maximum heat from the source with highest efficiency. The second, referred to as “comfort assist,” brings on 5 kW of resistance with the compressor at colder bins. For this study, the 5 kW was brought on starting with the last bin that did not require backup in the “best case” scenario. When the temperature becomes cold enough that the compressor plus the 5 kW of resistance is insufficient, additional resistance is brought on to meet the load. This strategy attempts to address the problems of “cold blow” (perceived cold air blowing out of the registers) at lower temperatures to improve comfort at the cost of some additional energy consumption. The third control strategy is the “low ambient cutout.” This strategy is also used to address the cold blow problem. In this strategy, the compressor is shut off whenever it is below a specified outdoor temperature, which effectively makes the equipment an electric furnace. Though the majority of installers do not use this strategy, it is a practice that is often implemented in colder climates (Kozak 2003). When the low 445

© 2004. American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). Published in ASHRAE Transactions, Vol. 110, Part 1. For personal use only. Additional reproduction, distribution, or transmission in either print or digital form is not permitted without ASHRAE’s prior written permission.

ambient cutout is used, the cutout temperature is often 30°F in Spokane. The final strategy considered here is referred to as “Plus 5.” As with the “comfort assist” and the “low ambient cutout,” this strategy is intended to avoid cold blow problems. In this strategy, 5 kW of resistance is wired to come on with the compressor whenever there is a heating call. As with the low ambient cutout, this strategy is not employed by all contractors, but it is not uncommon. For example, in a survey of 46 heat pumps in the greater Seattle area, about 30% had resistance come on whenever there was a call from the thermostat for heating (Davis 2001). The common theme in these control strategies is that cold blow problems result in occupant complaints and should be avoided. One of the primary goals of most contractors is to avoid customer callbacks (being called back to the house to fix a problem), and one of the most common reasons for callbacks is the perception of cold blow. This is often a problem with heat pumps because the temperature rise is often about 20°F (especially in colder bins). Given an entering temperature of 65°F to 70°F, this provides an outlet temperature of 85°F to 90°F. While this is warm enough to heat the house, it is colder than skin temperature, and thus the air coming out of the supply registers feels cool to the occupants. In order to avoid this problem, the registers need to be placed in locations where the exiting air will not be felt by the occupants. Common practice typically places registers under windows or as convenient to the installer, usually without regard to the final location of furniture or occupant living patterns. In the study by Davis (2001), there were some cases where an outdoor thermostat, which is intended to prevent the backup resistance from coming on above a certain temperature, had been disabled by a contractor, even if the device was still located within the outdoor unit. In addition, attempts to explain to homeowners that the supply air was warm enough to keep the house warm resulted in about an even split between homeowners who wanted to maximize their efficiency and those who preferred to settle for the higher bills in order to have warmer air come out of the registers. RESULTS The results are broken into several subtopics. The first discusses the impact of the various control strategies and duct losses on the performance of the heat pumps. The second section considers the improvement from the less efficient heat pump to the more efficient heat pump. The final section looks at how the different control strategies impact the duct efficiency specifically, as opposed to the system efficiency including both ducts and the equipment. Impact of Control Strategy and Duct Losses on Heat Pump Performance The primary results discussed in this paper are shown in Figures 2-5, which show heating season system efficiencies for a variety of combinations. The left half of each graph 446

shows the results for the case of no duct losses (e.g., all ducts inside the conditioned space), while the right side shows the results when duct losses are included. Figures 2 and 3 show the bin calculation results for the half-basement house in Seattle and Spokane, with Figures 4 and 5 showing the results for the crawlspace house. Figures 2 and 4 use the 8.2 HSPF heat pump, while Figures 3 and 5 use the 7.2 HSPF heat pump. In each graph, the line at a system efficiency of 1 corresponds to electric resistance heat with no duct losses (e.g., baseboard heat or electric furnace with all ducts within the conditioned space) and the higher horizontal line corresponds to the seasonal system efficiency specified by the HSPF rating for the heat pump in the graph (divide the HSPF rating by 3.413). For the 8.2 HSPF heat pump, this line is at 2.40, while for the 7.2 HSPF heat pump the line is at 2.11. For comparison, the performance of an electric furnace is shown in all graphs. All four of these figures show that Seattle is a very favorable climate for heat pumps. This is because Seattle is warmer than the typical Region IV site. This relatively mild climate causes the heat pump to operate primarily under conditions of higher compressor efficiency and minimizes the hours of backup electric resistance use. When there are no duct losses, the seasonal system efficiency is at or above the rating even with the comfort assist or low ambient cutout control strategies. In the best case, the system efficiency is about 6% to 9% higher for the 7.2 HSPF heat pump and about 10% to 14% higher for the 8.2 HSPF case. The small decrease for the comfort assist and low ambient cutout strategies are because there are very few hours below 30°F, so these strategies do not come into play very frequently. In fact, there is little difference between the comfort assist and low ambient cutout. The one control strategy of those evaluated that has a major impact in Seattle is the Plus 5 strategy, where 5 kW of resistance comes on with the compressor whenever heat is called for by the thermostat. At 47°F outside temperature, this increases the energy consumption of the 8.2 HSPF unit by about 160% while only increasing the output capacity by about 40%. The result is to eliminate nearly 70% of the potential benefit compared to baseboard heating, even without duct losses. In Spokane, even with its much colder climate, the heat pumps are at or near their rating for the best case with no duct losses. This indicates that Spokane is similar to the climate selected for the HSPF ratings and that, when operated properly and with minimal duct losses, heat pumps can be a good option in this location. The comfort assist strategy has a larger impact in Spokane because the additional backup resistance (beyond what is needed in the best case) is brought on much more frequently. The drop in seasonal system efficiency with the comfort assist strategy is about 8% in the half-basement house and about 3% in the crawlspace house when there are no duct losses. The low ambient cutout is much more problematic in Spokane than in Seattle due to the large number of hours below 30°F. In fact, unlike Seattle, the low ambient cutout in Spokane ASHRAE Transactions: Symposia

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Figure 2 Heating season system efficiency: half-basement house, higher efficiency heat pump.

Figure 3 Heating season system efficiency: half-basement house, lower efficiency heat pump.

Figure 4 Heating season system efficiency: crawlspace house, higher efficiency heat pump.

Figure 5 Heating season system efficiency: crawlspace house, lower efficiency heat pump.

bears more resemblance to the Plus 5 strategy than to the comfort assist strategy. This indicates that eliminating the compressor below 30°F in this climate has nearly the same effect as allowing the compressor below 30°F but having a minimum of 5 kW resistance at all times of operation. For the 8.2 HSPF heat pump, the low ambient cutout in Spokane eliminates about 60% of the potential benefit over electric baseboards, and for the 7.2 HSPF unit the drop is about 55%. The Plus 5 control strategy has a slightly larger but similar effect in Spokane compared to Seattle, typically eliminating just over 70% of the potential benefit over electric baseboards.

can make a major impact, despite the fact that the duct leakage and insulation levels were fairly typical of existing construction. In Seattle, the duct losses in the crawlspace house reduce the seasonal system efficiency by about 25% to 30%. For the Plus 5 control strategy with the 8.2 HSPF heat pump, this reduction in efficiency represents about 70% of the remaining potential benefit over electric baseboard heating and about 45% of the remaining potential benefit over the other strategies. For the 7.2 HSPF heat pump, these reductions increase to 80% and 49%, respectively. In Spokane, the duct losses make a bigger impact in the crawlspace case, in part because of the colder outdoor temperatures. The seasonal system efficiencies drop by 28% to 38% due to duct losses. In the Plus 5 case, the duct losses cause the seasonal system efficiency to drop below 1.0, meaning that it

Figures 2 and 3 show that the duct losses as described for the half-basement house do not cause much additional loss, about 10%. This is primarily because half the supply duct leakage was assumed to be inside the house and therefore not lost. Figures 4 and 5 show that the duct losses from the crawlspace ASHRAE Transactions: Symposia

447

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Figure 6 Seattle degree-hours, compressor capacity vs. load. Bars go with the left axis, lines go with the right axis.

Figure 7 Spokane degree-hours, compressor capacity vs. load. Bars go with the left axis, lines go with the right axis.

would have been more efficient to install electric baseboards to heat the home. In all cases, however, it is fair to point out that an electric furnace with the same duct losses would have a seasonal system efficiency lower than that from the Plus 5 case. Therefore, the heat pump is still using less energy than the electric furnace. Increased duct losses would make the heat pumps become less favorable compared to baseboards but would still be better than an electric furnace with the same ducts. Figures 6 and 7 provide another way to look at the effect of duct losses on the heat pump performance. These graphs show the compressor capacity of the 8.2 HSPF unit vs. outdoor temperature for the crawlspace house in Seattle and Spokane, respectively. In addition, the loads that must be met with and without ducts are shown. In these graphs, the duct losses are treated as an additional load that must be met by the heat pump. When the available compressor capacity is above the load, the compressor is able to meet the entire load without any backup assistance. When the compressor capacity drops below the load line, the area between the compressor capacity and the load is the amount of backup required to keep the house warm. The vertical line at 30°F in each graph reflects a common setting for the low ambient cutout, below which all heating is provided by the backup. These graphs show that, without duct losses, the heat pump is able to meet the load without backup down to about 7°F. However, with the ducts included, the heat pump is only able to meet the load without backup to about 22°F. In Seattle, this is not a major issue because the number of hours below 22°F is very small. In Spokane, however, there are a large number of hours below 22°F and, since the load is also highest at colder temperatures, these bins have even more importance. The area below the curve for the load with ducts and above both the compressor capacity and the load without ducts represents the additional backup resistance required as a result of duct losses. It is worth pointing out that, with ideal controls and no duct losses (such as by moving them into the condi-

tioned space), backup resistance would not be needed at all in the Seattle cases and only rarely in the Spokane cases using the 3.5 ton heat pump with an HSPF of 8.2.

448

Comparison Between Heat Pumps The previous section discussed factors that can cause the performance of a heat pump to be different from what is expected based on the stated ratings. However, that does not address the relative merits of one heat pump compared to another. This section looks at the extent to which the stated efficiency improvement from one heat pump to another holds regardless of duct losses and control strategies. Table 2 shows the ratio of the system efficiency for the 8.2 HSPF heat pump to that for the 7.2 HSPF heat pump for the various control strategies with and without duct losses. These values can be compared to an assumed 14% improvement based on the ratings themselves. This table shows that, when resistance is not being used to meet an unnecessarily large fraction of the load, the improvement from upgrading to the more efficient heat pump is actually often greater in these climates than the assumed improvement. This does not change significantly with the inclusion of duct losses. When the resistance is utilized to meet a large fraction of the load, even when it is not necessary, the relative improvement drops by about half to two-thirds. This is because, rather than having a system that behaves as a heat pump, the system is actually behaving as a hybrid heat pump/ electric resistance furnace, and the efficiency improvement is diluted by the electric resistance component. For Seattle, only the Plus 5 control strategy results in a realized improvement less than the assumed 14% improvement. For Spokane, the low ambient cutout also results in a greatly reduced improvement. This means that, as long as the equipment is installed such that the resistance truly only assists the compressor instead of displacing it, a consumer can ASHRAE Transactions: Symposia

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Table 2.

Ratio of System Efficiency from the 8.2 HSPF Heat Pump to the 7.2 HSPF Heat Pump Seattle

Spokane

Crawlspace

Half-Basement

Crawlspace

Half-Basement

Best Case

1.19

1.19

1.18

1.17

Comfort Assist

1.19

1.18

1.17

1.16

Low Amb. Cutout

1.15

1.16

1.07

1.07

Plus 5

1.05

1.06

1.05

1.06

Best Case w/Ducts

1.17

1.18

1.14

1.16

Comfort Assist w/Ducts

1.14

1.18

1.13

1.14

Low Amb. Cutout w/Ducts

1.15

1.16

1.06

1.07

Plus 5 w/Ducts

1.06

1.06

1.05

1.06

Table 3.

Effect of Control Strategy on Duct Efficiency1 Seattle

Spokane

Crawlspace

Half-Basement

Crawlspace

Half-Basement

Electric Furnace

0.80

0.91

0.78

0.90

Best Case

0.71

0.89

0.62

0.86

Comfort Assist

0.72

0.89

0.64

0.87

Low Amb. Cutout

0.73

0.89

0.71

0.89

Plus 5

0.76

0.90

0.70

0.89

1. An efficiency of 1.0 means no duct losses.

Table 4.

Effect of Heat Pump Sizing on Duct, Equipment, and System Efficiencies 3.5 ton

2.5 ton

Duct eff.

Equip. eff.

System eff.

Duct eff.

Equip. eff.

System eff.

Best Case

1

2.24

2.24

1

2.30

2.30

Plus 5

1

1.40

1.40

1

1.28

1.28

Best Case w/Ducts

0.72

2.28

1.63

0.67

2.34

1.56

Plus 5 w/Ducts

0.77

1.41

1.08

0.74

1.30

0.96

approximately expect to realize the relative improvement from selecting a more efficient unit. Effect of Control Strategy on Duct Efficiency The amount of heat that can potentially be delivered to a building is based on the difference in temperature between the delivered supply air and the house. If losses to a cold buffer space cause the air temperature in the ducts to drop toward the house temperature, then the duct efficiency approaches zero. Therefore, the lower the outlet temperature from the air handler, the more impact duct losses will have on the ability of the system to heat the building. Table 3 shows the duct efficiency for the various control strategies in each climate using the 8.2 HSPF heat pump. This table shows that, as the backup (electric resistance) component increases, the duct efficiency actually increases. This is ASHRAE Transactions: Symposia

because of the larger temperature rise across the equipment. The fact that the overall system efficiency decreases is due to the fact that the equipment efficiency decreases (with the introduction of electric resistance backup heat) by a larger percentage than the duct efficiency increases. This shows that the inclusion of duct losses narrows the gap between an electric furnace and a heat pump. Effect of Sizing on System Efficiency The sizing of the heat pump can have an impact on the system efficiency, both positive and negative. A heat pump with lower capacity will suffer smaller losses due to part-load operation. However, it will also require increased use of backup heat, with its lower efficiency, thus lowering the overall system efficiency at lower temperatures. Table 4 shows the effects of sizing on efficiency in the Seattle crawlspace house. 449

© 2004. American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). Published in ASHRAE Transactions, Vol. 110, Part 1. For personal use only. Additional reproduction, distribution, or transmission in either print or digital form is not permitted without ASHRAE’s prior written permission.

This investigation used two 7.2 HSPF heat pumps, one the 3.5 ton unit used elsewhere in this paper and the other a 2.5 ton unit. Only the best case and the Plus 5 cases were evaluated, both with and without duct losses. These results show that sizing did not make a large difference in any case. However, they also show that, while reducing the size does improve the equipment efficiency when there are no duct losses and the best case control strategy is used, the efficiency is always lower for the smaller unit for the other cases. This is because the increased use of resistance more than compensates for the reduction in part-load cycling losses. These results should not be used as a general assessment of the impact of changes in equipment sizing. The similarity of the numbers between the two heat pumps suggests that, in many cases, one or more of the comparisons may change sign for a different pair of heat pumps, different control strategies or duct losses, and different climates. CONCLUSIONS This study provides a number of insights into the performance of heat pumps in Pacific Northwest climates. Both the Seattle and Spokane climates show definite potential for the use of heat pumps, with Seattle being a very favorable location. When the systems are operated to maximize the benefit of the compressor, and when duct losses are small, the heat pumps perform as well as or better than advertised. However, the results also show that the benefits can be greatly reduced by a variety of factors, both intentional and unintentional. Control strategies that greatly increase the use of backup heat can seriously compromise the expected efficiency benefits. These strategies are not used by all contractors, but they are not uncommon. When they are used, it is done to avoid customer complaints about “cold blow,” often without the knowledge of the customer. Duct losses can also have a major impact on the efficiency of the system. Though this is true for any heating system with ducts outside the conditioned space, the effect is greater for heat pumps with similar duct conditions. The combination of duct losses and control strategies that utilize a lot of backup electric resistance heat can actually degrade the overall performance to a level that is worse than electric resistance without duct losses, such as electric baseboard heating (or an electric furnace with all ducts inside the conditioned space). Despite the problems that can be caused by installation problems, the relative benefit of a higher efficiency heat pump over a lower efficiency heat pump holds for these cases as long as the control strategy does not result in extensive additional use of backup heat. When this type of control strategy is used, the relative benefits are cut in about half. Due to higher delivery temperatures, the duct efficiency for combinations that use more backup heat is higher than for combinations that use less backup heat. This lessens the overall impact that the increased resistance has on the seasonal system efficiency. 450

Reducing the size of the heat pump results in an expected efficiency improvement but only when the compressor is considered. With the inclusion of duct losses and/or significant additional use of resistance heat in the control strategy, a substantially oversized heat pump provides a better seasonal system efficiency. This is contrary to the standard assumption that a heat pump sized to meet the load at a specific temperature, often 30°F, improves the overall efficiency due to a reduction in part-load losses. These results argue for significantly greater awareness on the part of heat pumps installers and customers. Installers need to better understand the ramifications of sub-optimal installations. Homeowners need to be better aware of possible installation problems. Program managers and regulators need to better understand the sorts of problems that can arise and design programs and codes to better ensure that heat pumps will provide the efficiency benefits of which they are capable. ACKNOWLEDGMENTS The authors thank the Regional Technical Forum of the Bonneville Power Administration for sponsoring this work, Tom Eckman project manager. REFERENCES ACCA. 1984. Acca Manual H: Heat Pump Systems: Principles and Applications, 2nd edition. Arlington, Va.: Air Conditioning Contractors of America. ARI. 1994. 1994 Standard for Unitary Air-Conditioning and Air-Source Heat Pump Equipment. Arlington, Va.: Air Conditioning & Refrigeration Institute. ASHRAE. 2003. ASHRAE Standard 152P, Method of Test for Determining the Steady-State and Seasonal Efficiencies of Residential Thermal Distribution Systems. Atlanta: American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. Davis, R. 2001. Field Performance of 50 Residential Air-toAir Heat Pumps. Final report for Puget Sound Energy. Ecotope, Inc., Seattle, WA. Environmental Protection Agency. 1998. EPA Energy Rating Process for ENERGY STAR® New Homes. Washington, D.C.: Environmental Protection Agency. Francisco, P. W., D. Baylon, and L. Palmiter. 2002. Estimating Deemed Savings from Residential Duct Insulation. Final report for the Bonneville Power Administration. Ecotope, Inc., Seattle, WA. Henderson, H., D. Parker, and J. Huang. 2000. Improving DOE-2’s RESYS Routine: User Defined Functions to Provide More Accurate Part Load Energy Use and Humidity Conditions. Proceedings of the 2000 ACEEE Summer Study on Energy Efficiency in Buildings, American Council for an Energy-Efficient Economy, Washington, D.C. International Code Council. 1998. International Energy Conservation Code, Sec. 201. Whitter, Calif.: International Council of Building Officials. ASHRAE Transactions: Symposia

© 2004. American Society of Heating, Refrigerating and Air-Conditioning Engineers, Inc. (www.ashrae.org). Published in ASHRAE Transactions, Vol. 110, Part 1. For personal use only. Additional reproduction, distribution, or transmission in either print or digital form is not permitted without ASHRAE’s prior written permission.

Kao, J. Y., W. J. Mulroy, and D. A. Didion. 1987. Field Performance of Three Residential Heat Pumps in the Heating Mode. National Bureau of Standards report NBSIR 87-3528, Gaithersburg, MD. Knebel, D.E. 1984. Simplified Energy Analysis Using the Modified Bin Method. American Society of Heating, Refrigerating, and Air-Conditioning Engineers, Inc., Atlanta, GA. Kozak, V. 2003. Inland HVAC Assoc., Spokane, WA. Personal communication to author. March 12. McQuiston, F.C. and J.D. Parker. 1988. Heating, Ventilating, and Air Conditioning: Analysis and Design, 3rd ed. John Wiley & Sons, New York, NY 1988. Palmiter, L. and P. W. Francisco. 1997. Development of a Practical Method for Estimating the Thermal Efficiency of Residential Forced-Air Distribution Systems. Electric Power Research Institute report TR-107744, Palo Alto, CA. Walker, I., J. Siegel, K. Brown, and M. Sherman. 1998. Saving tons at the register. Proceedings of the 1998 ACEEE Summer Study on Energy Efficiency in Buildings, American Council for an Energy-Efficient Economy, Washington, D.C. DISCUSSION Bert Phillips, P.Eng., Unies Ltd., Winnipeg, Canada: Regarding oversizing for heating may be beneficial. Heating contractors tend to oversize heating systems in any case so oversizing relative to contractor HL estimates results in gross oversizing. A heat pump sized to meet less than calculated design HL may in fact meet the actual peak heat loss. A heat pump designed to meet, say, 50% of the DHL will meet over 80% of annual heating load. This will significantly reduce capital cost for HP system. The savings could cover cost of a gas fireplace used for meeting the peaks in a house. Paul Francisco: One of the implications of the discussion of oversizing was that, to the extent that contractors oversize already, this is not necessarily a bad thing in heating mode. Cost/benefit is certainly a factor will come into play. In many cases, it may be that for little additional cost a larger heat pump can be used that will obviate the need for backup significantly. However, if the cost of a larger unit cannot be justified

ASHRAE Transactions: Symposia

by the improved efficiency due to infrequent occurrences of the colder temperatures, then the financial resources could be put to more effective use. Also, once the compressor can meet the load at the coldest condition, no more oversizing is appropriate, since the primary change after that point will be to increase cycling losses. Ron Judkoff, Director Center for Buildings and Thermal Systems, Golden, Colo.: Unitary heat pump cabinets are more poorly insulated and leakier than the buildings they are attached to, thus exacerbating poor performance in cold climates. Dave Baylon, Ecotope, Inc., Seattle, Wash.: You mentioned that the duct efficiency associated with heat pumps is worse than the duct efficiency with an electric or gas furnace. Could you clarify this comment—why would similar duct have lower efficiency? Francisco: One primary reason why this is so is because of the lower temperature rise across the heat pump coil. Because the temperature difference between the supply air and the house is usually significantly lower than the temperature rise for a furnace, a loss of heat due to conduction or leakage losses is a greater fraction of the total possible useful heat. Another factor that effects the overall efficiency is the increased use of backup heating. This backup heating will increase the temperature rise, such that the efficiency through the ducts may actually be improved, but the overall efficiency will be greatly affected because of the lower efficiency of the backup heat relative to the compressor. Jim Crawford, Director of Regulatory Affairs, Trane, Tyler, Tex.: (1) In the final Q&A, a representative of NREL made an observation to the effect that packed heat pump ratings are optimistic due to cabinet cooling by ambient air. In this context, it is important to note that performance tests of such products per ARI-210/240 are performed with the cabinet in a simulated ambient. Thus, the effect mentioned is factored into the test. (2) The effect of ambient temperatures on performance of single package (or “packaged”) heat pumps is reflected in a slightly less stringent standard under the current NAECA rules.

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