Retrofitting Combined Space 

and Water Heating Systems: 

Laboratory Tests

B. Schoenbauer, D. Bohac, P. Huelman, 

R. Olson, and M. Hewett

NorthernSTAR

October 2012
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iii

Retrofitting Combined Space and Water Heating Systems:

Laboratory Tests

Prepared for:

The National Renewable Energy Laboratory 

On behalf of the U.S. Department of Energy’s Building America Program

Office of Energy Efficiency and Renewable Energy

15013 Denver West Parkway

Golden, CO 80401

NREL Contract No. DE-AC36-08GO28308

Prepared by:

B. Schoenbauer, D. Bohac, P. Huelman,

R. Olsen, and M. Hewett

NorthernSTAR Building America Partnership

322 Kaufert Lab

2004 Folwell Avenue

Saint Paul, MN 55108

NREL Technical Monitor: Stacey Rothgeb

Prepared under Subcontract No. KNDJ-0-40338-00

October 2012
iv

Acknowledgments

This report was prepared for the U.S. Department of Energy Building America program. Funding 

for this work was provided by the National Renewable Energy Laboratory under contract KNDJ0-40338-00. Additional funds to support this work were provided by the Weatherization 

Assistance Program through a Sustainable Energy Resources for Consumers grant, the Center for 

Energy and Environment, and the Energy Conservatory. Additional support through loaned or 

donated equipment was provided from Rinnai, A.O. Smith, American Water Heaters, Eternal, 

Heat Transfer Products, Navien America, Enerzone, Lifebreath, NuAire, and Advanced 

Distributor Products.

The primary authors were Ben Schoenbauer, Dave Bohac, and Martha Hewett, Center for Energy 

and Environment, and Jake McAlpine, Sustainable Resource Center. 

These authors would like to acknowledge the contributions and support of Becky Olson, 

Sustainable Resource Center, for project management and cost analysis; Pat Huelman, 

University of Minnesota, for expertise and role as an advisor to the project; Alex Haynor, 

Corinne Wichser, and Josh Novacheck, Center for Energy and Environment for work on field 

installations and data analysis; Anna Jursik, Center for Energy and Environment for report 

editing; Gary Nelson, The Energy Conservatory for donating space for the laboratory and his 

time as an advisor; Gary Klein, Affiliated International Management, for his time and expertise; 

Louise Goldberg, University of Minnesota–BBE, for the BEopt modeling and analysis; and Tom 

Schirber, University of Minnesota, for coordinating and editing reports.
v

Contents

Acknowledgments ..................................................................................................................................... iv

List of Figures ............................................................................................................................................ vi

List of Tables ............................................................................................................................................. vii

Definitions................................................................................................................................................. viii

Executive Summary................................................................................................................................... ix

Introduction.............................................................................................................................. ix

Experiment............................................................................................................................... ix

Analysis.................................................................................................................................... ix

Results...................................................................................................................................... ix

Conclusions.............................................................................................................................. xi

1 Introduction........................................................................................................................................... 1

1.1 Background..........................................................................................................................2

1.2 Relevance to Building America’s Goals..............................................................................3

1.3 Cost Effectiveness................................................................................................................3

1.4 Tradeoffs and Other Benefits...............................................................................................4

1.5 The Role of Combi Systems in Low Income Weatherization .............................................4

2 Experiment ............................................................................................................................................ 5

2.1 Research Questions..............................................................................................................5

2.2 Technical Approach.............................................................................................................5

2.3 Measurements......................................................................................................................6

2.4 Combi Equipment and Test Instrumentation .......................................................................8

3 Analysis ............................................................................................................................................... 12

3.1 Standby Loss and Steady-State Efficiency Analysis.........................................................12

3.2 Air Handler Capacity Analysis..........................................................................................13

3.3 Full System Analysis .........................................................................................................13

3.4 Error Analysis....................................................................................................................13

4 Results................................................................................................................................................. 15

4.1 Standby Loss......................................................................................................................15

4.2 Steady-State Efficiency......................................................................................................18

4.3 Air Handler Capacity .........................................................................................................19

4.4 Full System ........................................................................................................................23

4.4.1 Maximum Capacity................................................................................................26

4.4.2 Response to Overlapping Events...........................................................................27

4.4.3 Domestic Hot Water Supply Temperature.............................................................29

4.5 Installation Process and Specifications..............................................................................32

5 Discussion........................................................................................................................................... 35

5.1 Air Handler Improvement..................................................................................................35

5.2 Heating Plant Improvement ...............................................................................................35

6 Conclusions ........................................................................................................................................ 37

7 References .......................................................................................................................................... 38
vi

List of Figures

Figure 1. Energy consumption during standby mode operation at a set point of 140°F, except 

Boiler 2, which has a fixed internal hot water set point of 156°F .................................................... x

Figure 2. Heating plant steady-state space heating efficiency over a range of return water 

temperatures........................................................................................................................................ xi

Figure 3. Diagram of a combined space and water heating system...................................................... 1

Figure 4. Test setup for the heating plant with a simulated AH ............................................................. 7

Figure 5. Test setup for the AH testing..................................................................................................... 8

Figure 6. Photos of test laboratory with combi system installations and venting terminations...... 11

Figure 7. Energy consumption during standby mode at a set point of 140°F.................................... 16

Figure 8. Infrared and visual images of Boiler 4 in standby operation ............................................... 16

Figure 9. Infrared and visual images of TWH 1 in standby operation ................................................. 17

Figure 10. Infrared and visual images of TWH 2 in standby operation ............................................... 17

Figure 11. Heating plant steady-state efficiency over a range of return water temperatures........... 18

Figure 12. Heating plant steady-state efficiency for 60 kBtu/h heating load ...................................... 19

Figure 13. AH7 performance mapping test of supply air temperature ................................................ 20

Figure 14. AH7 performance mapping test of return water temperature ............................................ 21

Figure 15. AH7 performance mapping test of space heating capacity ............................................... 21

Figure 16. AH6 measured and manufacturer specified water temperature difference...................... 23

Figure 17. Maximum capacity test profile for full system tests ........................................................... 24

Figure 18. DHW interrupt load profile for full system tests .................................................................. 24

Figure 19. Low use load profile for full system tests ............................................................................ 25

Figure 20. Difference in supply temperature from the start to end of a two shower and space 

heating demand .................................................................................................................................. 27

Figure 21. DHW supply temperature response when space heating demand was initiated............. 28

Figure 22. DHW temperature changes before and after the mixing valve for TWH 1 ........................ 29

Figure 23. DHW supply temperature profile for a short 1.2-gpm draw................................................ 30

Figure 24. DHW supply temperature profile for TWH 2......................................................................... 31

Figure 25. Typical outdoor setback temperature control curve........................................................... 34

Figure 26. Typical combi system installation with a boiler .................................................................. 36

Unless otherwise indicated, all figures were created by NorthernSTAR.
vii

List of Tables

Table 1. Laboratory Instrumentation......................................................................................................... 9

Table 2. Laboratory Tested Combi Heating Plants.................................................................................. 9

Table 3. Laboratory Tested Hydronic AHs ............................................................................................. 10

Table 4. Standby Site Energy Consumption at a Set Point Temperature of 140°F ............................ 15

Table 5. Sizing Chart for AHs for a 140°F Supply Water Temperature ................................................ 22

Table 6. Full System Test Load Profiles (Btu/h)..................................................................................... 25

Table 7. DHW Draw Characteristics for the Low Use Load Profile ...................................................... 29

Table 8. DHW Performance of Combi Systems Under Small and/or Short Draw Conditions........... 31

Table 9. Combi System Sizing Chart....................................................................................................... 33

Unless otherwise indicated, all tables were created by NorthernSTAR.
viii

Definitions

AFUE Annual fuel utilization efficiency

AH Air handler

BEoptTM Building Energy Optimization (software program)

cfm Cubic feet per minute

Combi

DHW

Combined space and water heating system

Domestic hot water

EF Energy factor

gpm Gallons per minute

TWH Tankless water heater

WH Water heater
ix

Executive Summary

Introduction

Better insulation and tighter envelopes reduce space heating loads for new and existing homes.

For many homes, decreased space heating loads enable a single heating plant to provide space 

and domestic water heating loads. These systems, called integrated appliances or combination 

(combi) systems, can also eliminate safety issues associated with natural draft appliances by 

using one common sealed combustion vent.

The research will use combi systems with water heater (WH) or boiler heating plants teamed 

with forced air distribution space heating systems. In each house, the existing furnace will be 

replaced with a hydronic air handler (AH) that includes an AH, a water coil, and a water pump to 

circulate water between the heating plant and the coil. The tank type WH will be replaced by 

either a WH or a boiler with a separate circuit for domestic hot water (DHW). Various options 

for DHW priority, DHW tempering, and heating plant temperature set point control will be 

considered.

Experiment

Initial bids indicated that local mechanical contractors had only limited experience in the design 

and installation of high efficiency combi systems. The NorthernSTAR combi laboratory was 

created to identify proper system components, designs, operating parameters, and installation 

procedures to ensure field installed systems are highly efficient. The laboratory also provided a 

place for contractors, utility representatives, weatherization agents, and codes officials to view 

the systems and become familiar with their installation. Nine heating plants were installed in the 

laboratory space. Four condensing combi boilers with heating loops for space and domestic 

water heating (Boiler 1, Boiler 2, Boiler 3, and Boiler 4) were tested Two (Boiler 2 and Boiler 4)

had internal storage for hot water . Three condensing storage tank type WHs were installed 

(Tank 1, Tank 2, and Tank 3): one condensing tankless water heater (TWH 1) and a condensing 

hybrid WH that has a 199,000 Btu/h modulating burner and 2 gal of storage capacity (TWH 2).

Analysis

Laboratory tests—heating plant standby mode energy use, space heating steady-state efficiency, 

AH capacity, and full system output capacity—were used to optimize combi system efficiency.

Data loggers were used with water flow rate and water temperature differences to measure 

system energy output. Gas meters and electric watt transducers were used to measure energy and 

power input at 1-s intervals. The data were then analyzed to determine the best operating 

parameters and system components. Hydronic AH steady-state performance measurements 

determined output capacities that provided acceptable return water and supply air temperatures.

Heating plant capacity results were used to develop algorithms to determine whether a system 

could meet DHW and space heating loads. Combined space heating and DHW load profiles were 

used to evaluate transient performance.

Results

Standby tests showed that some heating plants had much higher loss rates than others. Figure 1

shows idle energy consumption at a set point of 140°F. Heating plants with no hot water storage 

(Boiler 1, Boiler 3, and TWH 1) had no gas use; the gas use for the units with storage varied 

from 29 to 103 therms/yr. The standby consumption of the three tank type WHs was about equal 
x

to or lower than that of the tankless water heater (TWH 2) and boilers with storage (Boiler 2 and 

Boiler 4). TWH 2, Boiler 2, and Boiler 4 had standby losses greater than desirable considering 

those units had notably smaller storage volumes than the tank type heaters. For these systems, 

the level of insulation appeared to have more effect than the storage size on idle losses.

Figure 1. Energy consumption during standby mode operation at a set point of 140°F, 

except Boiler 2, which has a fixed internal hot water set point of 156°F

Figure 2 shows the steady-state efficiency of each heating plant for a range of return water 

temperatures, 90°–120°F, which resulted in heating load of 20,000 to 100,000 Btu/h. (A set point 

temperature of 130°F and a flow rate of 4 gpm were used for all tests.) The heating plants had 

similar steady-state efficiencies with a low of 86.1% for Boiler 3 and high of 90.6% for TWH 2.

Given that the uncertainty of the efficiencies was 1.5 percentage points, the difference in steadystate efficiency was significant only for the highest and lowest efficiency systems. The average 

efficiency of the four boilers was 1.8 percentage points lower than the average for the TWHs and 

1.9 percentage points lower than the tank type WH average. However, the difference in average 

efficiency between the three system types was within the expected measurement error.

0

200

400

600

800

1000

1200

1400

Boiler 1 Boiler 2 

(156)

Boiler 3 Boiler 4 TWH 1 THW 2 Tank 1 Tank 2 Tank 3

Energy Consumption During Stand by [Btu/hr]

Gas Electric
xi

Figure 2. Heating plant steady-state space heating efficiency over a range of 

return water temperatures

AH capacity testing determined the available heat output rate for each AH. AH capacity was 

determined by requiring the water temperature leaving the AH to be ≤105°F. The delivered air 

temperature was required to be ≥110°F to prevent occupant discomfort. These requirements 

greatly restricted the available capacities from those specified by the manufacturers. 

Conclusions

The highest system efficiencies were achieved by minimizing the water temperatures returning 

from the hydronic AH. Heating loads were determined for each site. These loads were used to 

select the water flow rates and airflow rates that would result in low return water temperatures 

and meetg the necessary load with an acceptable air temperature. Laboratory tests showed that 

the heating plant steady-state efficiency decreased with increasing return water temperature. The 

decrease in efficiency became more significant as the return temperature increased above 110°F.

Laboratory testing verified that systems were capable of meeting heating loads up to 50,000 

Btu/h with acceptable return water temperatures and supply air temperatures. These designs 

provide steady-state space heating efficiencies >85%.

System design and sizing information developed in the laboratory is currently being used to 

optimize systems for the 300 site implementation project. The laboratory work has been 

invaluable to this process. The installers and program managers thus have confidence that these 

systems will meet the needs of the homeowners.

75%

80%

85%

90%

95%

100%

70 80 90 100 110 120 130

Steady-State Space Heating Efficiency

Return Water Temperature, °F

TANK 1 TANK 2 TANK 3 BOIL 1 BOIL 2 BOIL 3 BOIL 4 TWH 1 TWH 2

Set Temp = 130F, Flow = 4 
xii

The performance of combi systems is limited by currently available equipment. Most hydronic 

AHs were not designed to produce lower return water temperatures necessary for combi systems 

with a condensing heating plant. Several manufacturers, at least one in direct reaction to the 

findings from this project, have begun to improve combi equipment. AHs with improved heat 

transfer performance will allow for lower and lower return water temperatures. Variable flow 

rate water pumps and fans, along with the necessary control capabilities, have significant 

potential to provide a greater range in heating output and to provide low return water 

temperatures. Also, the manufacturer required that the primary secondary loop configuration for 

combi systems using boilers significantly increases return water temperatures, reducing system 

efficiency. Modifications to this configuration could improve system efficiency as much as 10%.
1

1 Introduction

Better insulation and tighter envelopes are reducing space heating loads for new and existing 

homes. For many homes, decreased space heating loads enable a single heating plant to provide 

space and domestic water heating loads. These systems, called dual integrated appliances or 

combination (combi) systems, can also eliminate safety issues associated with natural draft 

appliances because they use one common sealed combustion vent. Figure 3 shows a typical 

combi system set up, with a heating plant and a hydronic air handler (AH). During a hot water 

draw the heating plant operates like a water heater (WH): cold water comes into the storage 

volume as hot water leaves (for units with storage). In systems without storage, cold water enters 

and is heated as necessary. During a space heating event, hot water leaves the heating plant,

passes through the coil in the AH, and transfers heat into the airflow. The cooler water then 

leaves the AH and flows back to the heating plant. The figure shows an open loop system where 

potable water is circulated to the AH for space heating. Closed loop systems use a heat 

exchanger between the heating loop in the heating plant and the AH.

Figure 3. Diagram of a combined space and water heating system

The research will use combi systems with condensing WH or boiler heating plants teamed with 

forced air distribution space heating systems. In each house, the furnace will be replaced with a 

hydronic AH that includes an AH, water coil, and water pump to circulate water between the 

heating plant and the coil. The tank type WH will be replaced by either a WH or a boiler with a 

separate circuit for domestic hot water (DHW). Various options for DHW priority, DHW 

tempering, and heating plant temperature set-point control will be considered.

This project will use several types of condensing heating plants to characterize the installed 

performance of combi systems. The primary objective will be to document energy savings, and 

Heating 

Plant

DHW Loads

Supply to AH

Return from AH

Hydronic 

Air 

Handler

Hot Space 

Heating Air Cold Water Mains
2

the monitoring will include key system parameters such as return water temperature to better 

understand variations from expected performance and identify improved system designs. The 

project will also determine the installation costs and potential cost reductions with widespread 

implementation.

1.1 Background

Historically, mechanical contractors have custom engineered and pieced together combi systems 

in the field. They focused on assembling functional systems and had neither the time nor the 

means to evaluate or optimize the systems’ energy efficiency. As high efficiency condensing 

WHs and boilers gain a larger share of the residential market, there is greater potential to use 

these systems to improve the efficiency of space heating and DHW loads.

Research is needed to address several outstanding questions about combi systems, such as:

• What is the actual installed energy savings of an optimized combi system with a 

condensing heating plant? 

• What are the installation costs and paybacks of these systems?

• Can contractor familiarity and experience with combi systems reduce installation prices?

The concept of a single heating plant to supply space and water heating has been around for 

many years. Bohac et al. (1995) installed and monitored combis in small commercial and 

multifamily buildings in 1990. These systems used a natural draft storage WH to generate heat.

The 1.5 years of monitored operation demonstrated that these systems could be reliably installed,

perform without failure, and save energy. The combi systems in this project had annual fuel 

utilization efficiency (AFUE) ratings of at least 78% and replaced natural draft WHs with 

efficiencies in the 50% range and furnaces with AFUEs around 60%. The study found that these 

systems saved an average of 24% in energy use annually.

Noncondensing natural draft WHs were used in earlier combi systems. Condensing heating 

plants substantially increase the energy savings potential of combi systems. Laboratory testing 

(Thomas 2010) demonstrated that when combi systems replace mechanical equipment in modern 

homes, they must use condensing heating plants to achieve similar or improved energy 

performance.

Combi systems using high efficiency heating plants are relatively new. Several laboratory test 

and field installations identified potential problems. Laboratory tests by Brookhaven National 

Laboratory (Butcher 2011) showed that the manufacturer-specified plumbing configuration with 

a primary and secondary loop made it difficult to achieve the high efficiency potential of 

condensing combi boilers. A field installation by the New York State Energy Research and 

Development Authority (Rudd 2010) examined durability issues for systems using tankless water 

heaters (TWHs). The study assessed problems with hard water, scaling, and short cycling. Water 

residues accumulated on the TWH inlet filter and eventually prevented the heater from 

activating. The study installed an industrial strainer on the inlet water line, which prevented WH

failure and reduced the maintenance interval to an annual filter change.
3

A large sample field study will help determine how these systems work in the real world, and 

will assess the actual installed efficiency and performance of combi systems. These field tests 

will use current high efficiency products tested and optimized in the laboratory to determine the 

actual energy savings of well-designed combi systems. This report provides results from the 

laboratory tests; a later report will cover results from the field study.

1.2 Relevance to Building America’s Goals

Combi systems have potential to significantly reduce home energy use. An optimized combi

system can provide space and water heating with a 95% efficient heating plant, compared to a 

minimum efficiency 78% AFUE furnace and a 0.591 energy factor (EF) WH. These metrics 

cannot be compared directly; detailed laboratory and field monitoring is necessary to determine 

actual savings. The large difference between the efficiency numbers shows the potential for 

savings. Removing a naturally drafted WH also allows the home to be more airtight without 

causing combustion safety issues, and can eliminate combustion makeup air. These two 

measures can further improve the energy performance of a home.

Combi systems are feasible in most climate regions. In colder climates the application may not 

be possible in larger homes with poorly insulated and leaky envelopes. Currently available 

equipment should target homes with space heating loads smaller than 60,000 Btu/h.

The implementation phase of this project will install 300 combi systems in Minnesota homes. At 

completion, the contractors should be ready to install the systems across the state. This project 

will also develop installation guidelines and specifications to increase the success of installation 

decisions in all climates.

1.3 Cost Effectiveness

The installed cost of a high efficiency combi system may be lower than that of a similar 

efficiency separate furnace and WH. In a retrofit application, a homeowner can expect to pay 

approximately $4,000 for a high efficiency (90%–95% AFUE) furnace and $4,000 for a high 

efficiency (0.80–0.95 EF) WH.

2 In the Minneapolis area, contractors with limited experience 

with combi systems currently bid a high efficiency system for $6,000–$9,000. A large number of 

installations are expected to reduce costs, much as the cost of high efficiency water heaters

decreased as installations increased over the past few years.

A preliminary EnergyPlus analysis of a high efficiency boiler used for space and DHW heating 

estimated that natural gas use would be reduced by 12% and source energy by 7% compared to 

an 80% AFUE furnace and 0.55 EF WH. However, EnergyPlus is not easily adapted and may 

not properly model high efficiency combi system performance. EnergyPlus currently does not 

have models for performance of heating plants or combi systems. For example, it does not 

provide information on variations of efficiency with return water temperature. The test 

laboratory measurements for individual combi components are expected to provide performance 

data necessary for improved EnergyPlus models. Development of EnergyPlus combi system 

models will be completed in cooperation with National Renewable Energy Laboratory staff, with 

 1 For a gas-fired water heater with a storage volume of 40 gal. 2 High efficiency furnace and water heater numbers were determined through discussions with local installers.
4

the ultimate goal of incorporating combi system models into Building Energy Optimization 

(BEoptTM) software.

1.4 Tradeoffs and Other Benefits

The combi system provides several secondary benefits. It replaces a separate furnace and WH

with a single boiler or WH. This reduces the number of gas lines and exhaust vents from two to 

one, and a wall-mounted heating plant may reduce the equipment footprint. A combi system can 

allow for a more versatile mechanical room. In new construction, this may allow the mechanical 

equipment to be placed closer to the end uses, reducing delivery losses and hot water wait times.

A single high efficiency burner also has combustion safety and venting benefits. The high 

efficiency combi heating plants have power vent or direct vent exhaust systems. These eliminate 

combustion safety issues that arise as homes become better insulated and tighter with unbalanced 

ventilation. Sealed combustion burners may provide an additional energy benefit by eliminating 

combustion makeup air openings and possibly sealing a common chimney vent.

The single heating plant system also presents tradeoffs. The installation becomes more complex, 

which often requires onsite engineering and optimization to achieve maximum performance.

Also, many contractors are unfamiliar with these systems.

The implementation phase of the project was designed to ensure that the installed combi systems 

include a warranty and expected durability at least equal to the alternatives. Combi systems also 

have a single heating appliance that is expected to require less annual maintenance than a 

separate furnace and WH.

As part of the test laboratory phase of the project, properly installed combi systems have been

demonstrated to code officials and contractors who are bidding on the 300 field installations. The 

demonstrations allowed code officials to identify potential code concerns, such as water 

stagnation in potable rated AHs, and recommend acceptable solutions. Codes officials are thus 

more willing to accept this newer technology, and issues requiring attention during the 

installation and inspection process have been reduced.

1.5 The Role of Combi Systems in Low Income Weatherization

Advances in insulation, air sealing, and mechanical equipment have enabled low income 

weatherization to reduce heating loads and significantly improve home performance in 

Minnesota. As home airtightness has been reduced and exhaust ventilation is installed, there has 

been a greater concern with combustion gas spillage of natural draft WHs and other appliances.

In many homes, weatherization agencies have been forced to spend about $1,500 from a health 

and safety budget to install a power vented WH typically with an EF of 0.60 and only a small 

energy improvement. A combi system with a condensing boiler or WH eliminates the 

combustion spillage issue and provides higher efficiency DHW.
5

2 Experiment

Initial bids indicated that local mechanical contractors had only limited experience in the design 

and installation of high efficiency combi systems. The NorthernSTAR combi laboratory was 

created to identify proper system components, designs, operating parameters, and installation 

procedures to ensure high efficiency of field installed systems. The laboratory also provided a 

place for contractors, utility representatives, weatherization agents, and codes officials to see the 

systems and become familiar with their installation.

2.1 Research Questions

The laboratory tests address the following research questions:

• What equipment and design characteristics affect system performance and how can 

systems be optimized?

• What are the maximum and minimum heating loads that a combi system can meet 

efficiently and effectively?

• What are the minimum performance criteria, installation specifications and quality 

control methods that must be achieved to ensure proper performance and expected 

efficiency? What trouble areas of the installation can be addressed through these criteria 

and quality control methods?

• How can these systems benefit from improved combi products (boilers, WHs, hydronic 

AHs, pumps, controls, etc.)?

2.2 Technical Approach

This project designed and optimized combi systems for a large-scale implementation project.

The Sustainable Resource Center, a low income weatherization provider in Minnesota, received 

a Sustainable Energy Resources for Consumers grant to install 300 combined space and water 

heating systems in homes participating in the State of Minnesota Low-Income Weatherization 

Assistance Program. The grant includes a monthly utility bill analysis of pre- and postinstallation analysis of all 300 homes and of daily average space and DHW heating energy use 

for 20 of the homes.

Laboratory tests were conducted on a variety of complete combi systems and individual 

components before field installations started. Hydronic AH steady-state performance 

measurements determined output capacities that provided acceptable return water and supply air 

temperatures. Heating plant capacity results were used to develop algorithms to determine 

whether a system could meet DHW and space heating loads. Multiple systems were configured 

in the laboratory and experienced contractors reviewed initial designs to provide 

recommendations to improve performance, reliability, ease of installation, and cost. Twenty of 

the field systems will be extensively monitored to characterize their performance under real load 

conditions. The monitoring will determine overall system efficiency and the effects of many 

operational variables.
6

2.3 Measurements

The laboratory testing phase of the project documented the performance of currently available 

components and developed recommendations for optimized combi system designs. This required

a series of tests on the heating plants, hydronic AHs, and fully assembled systems:

• Heating plant

o Idle period energy use

o Steady-state space heating efficiency for varying return water temperature

o DHW transient supply water temperature

• Hydronic AH

o Steady-state heat output for varying water and air flow rates

• Full system

o Maximum capacity

o Cyclic tests.

The idle tests were conducted by recording gas and electricity consumption when the heating 

plant had no DHW use or space heating load. The duration of the test included at least five

burners on and recovery periods. All heating plants were tested at the same temperature set point 

and ambient conditions. Measured data were used for quantitative analysis of idle performance; 

infrared thermography was used to help identify sections of the heating plant with significant 

heat loss. 

The heating plant steady-state space heating efficiency tests were conducted using a “work 

horse” heating plant to supply hot water at a set flow rate and temperature to simulate the water 

conditions returning from a hydronic AH (see Figure 4). System efficiencies were computed by 

measuring the energy into the heating plant, in natural gas and electricity, and the energy output 

in hot water. Measurements were conducted for return water temperatures of 80°–120°F to 

characterize the efficiency degradation with higher return temperatures. This determined an 

optimized water circulation flow rate and heat output sizing of the hydronic AH. Laboratory tests 

were used to analyze the transient performance of the combi systems. These tests were not used 

to determine annual combi system efficiencies, but to examine delivery capabilities. Water 

temperatures were tracked during a sudden increase in DHW flow to determine hot water 

delivery times. The tests also examined transient supply water temperature under three simulated 

conditions: (1) a space heating event is interrupted by a DHW event (e.g., a shower); (2) a 

shower is interrupted by a second shower; and (3) a shower is interrupted by a space heating 

event.
7

Figure 4. Test setup for the heating plant with a simulated AH

The hydronic AH tests characterized the performance of units in conditions within and outside 

the range of manufacturer specifications. Manufacturers often do not specify performance for the 

lower water circulation flow rates necessary for optimized heating plant efficiency. The tests 

measured the heat output rate, supply air temperature, return water temperature, and electricity

consumption for controlled variations in airflow rate, water flow rate, water supply temperature, 

and air inlet temperature (see Figure 5). Space heating output capacity was computed from the 

water flow rate of the space heating loop and the temperature drop over the AH. Air 

temperatures were measured with thermocouples: a three thermocouple array on the supply side 

and a single probe in the return duct. Airflow rates were measured with a TrueFlow AH flow 

meter in the return duct (Energy Conservatory 2008). Air temperatures entering the AH were 

used to verify consistent conditions in the laboratory. Supply air temperatures were used to 

determine whether the delivered air would provide acceptable occupant comfort, but were not 

used for the energy output analysis. This methodology uses air temperature and flow rates only 

to determine occupant comfort, which did not require measurements to be highly accurate. Air 

temperatures were measured with thermocouples with ± 1°F accuracy compared to the water 

temperature measurements that were made with ± 0.1°F accuracy. 
8 

Figure 5. Test setup for the AH testing

Combi systems were operated using several hourly load patterns to determine cyclical 

performance and maximum capacity. Several DHW and space heating load profiles were

developed to evaluate the warmup and post purge portions of the heating plant firing cycle. 

Additionally, the tests required the combi system to simultaneously supply heat for two showers 

and a space heat cycle. The load profiles for these tests are included in the results section. 

2.4 Combi Equipment and Test Instrumentation

The test laboratory was equipped with a comprehensive and accurate monitoring system utilizing 

high-precision instruments (see Table 1). A Campbell Scientific model CR-3000 data logger was

programmed to measure instrument outputs and record processed data at specified intervals. A 

propagation of errors method using the uncertainties of individual instruments for typical 

operating conditions computed an uncertainty of 2% for the calculated hot water energy output 

and an uncertainty of 2.5% for the system efficiency. Three identical sets of monitoring systems 

were used for tests on nine heating plants and five AHs (see Table 2 and Table 3). Figure 6 

includes photos of some of the systems installed in the laboratory. All equipment was installed to 

meet manufacturer’s specifications. Heating plant vents were run as two pipe systems meeting in 

a concentric termination kit at the roof. One unit required concentric venting for the full vent run. 

Vent lengths were about 30 ft, well within the specifications for all manufacturers.
9 

Table 1. Laboratory Instrumentation

Measurement Sensor Type Resolution Precision Range

Water Volume 

Flow Rate

Nutating disk flow 

meter

198.4 pulses/

gal 2% of reading 0.5–25 gpm*

Natural Gas 

Volume

Diaphragm meter, 

with pulse output 40 pulses/ft3 0.3% of reading 0–250 cfm

Water 

Temperatures

Matched pair of 

immersion resistance 

temperature 

detectors

0.002°F 

@140°F

1/10 DIN:

0.03°F @32°F

–148° to 

752°F

Electric

Energy Watt transducer 0.02 W 0.2% of reading 0–1000 W

Air 

Temperature Thermocouple array 0.03°F 

@140°F

Greater of 1.8°F 

and 0.75% of 

reading

–454° to 

725°F

* The meter measures flow rates <0.5 gpm, but the precision decreases for flow rates outside the specified range.

Table 2. Laboratory Tested Combi Heating Plants

Equipment 

ID Equipment Type Manufacturer Model Input (kBtu/h)

Storage 

(gal)

BOIL 1 Condensing boiler Rinnai E75C 17–75 0

BOIL 2 Condensing boiler Rinnai Q175C 35–175 7

BOIL 3 Condensing boiler Navien Combi Boiler 20–175 0

BOIL 4 Condensing boiler Triangle Tube Prestige 30–110 12

TWH 1 Condensing TWH Rinnai 98Lsi 9.5–199 0

TWH 2 Hybrid condensing

TWH Grand Hall Eternal 31–199 2 

TANK 1 Condensing storage 

WH A.O. Smith Vertex 100 50 

TANK 2 Condensing storage 

WH HTP Phoenix 35–100 55 

TANK 3 Condensing storage 

WH American Polaris 100 34 
10

Table 3. Laboratory Tested Hydronic AHs

Equipment 

ID Equipment Type Manufacturer Model

AH 1 Hydronic AH Rinnai 37AHB060

AH 2 Hydronic AH Rinnai 37AHB075

AH 3 Hydronic AH ADP BVR00031s4p3

AH 4 Hydronic AH Nu-Air en7030i

AH 5 Hydronic AH Enerzone xah70vs-x13-pt-4row

AH 6 Hydronic AH First Company 8VMR

AH 7 Hydronic AH with heat recovery 

ventilator Lifebreath CFA-U-S4A-24-E16

AH 8 Hydronic AH with heat recovery 

ventilator NuAir EN712E

AH 9 Nonpotable hydronic AH Lennox cbwmv-36c-090-1

Table 2 lists the nine heating plants installed in the laboratory. Of these, four were classified as 

boilers, three as tank type WHs and two as TWHs. Each of the boilers was a combi boiler with 

both space heating and domestic water heating loops. Two (Boiler 1 and Boiler 3) were low mass 

boilers and contained no internal storage. TWH 1 was a traditional condensing; TWH 2 was a 

hybrid between a TWH and a storage WH. This unit had a large burner (199,000 Btu/h at full 

capacity) and a small (2-gal) storage tank.
11

Figure 6. Photos of test laboratory with combi system installations and venting terminations
12 

3 Analysis

3.1 Standby Loss and Steady-State Efficiency Analysis

The laboratory heating plant, hydronic AH, and full system test procedures are described in 

Section 2.3. The heating plant idle energy use was computed from the average rate of natural gas 

and electricity use when the system went through four to five burner cycles necessary to keep the 

unit at the operating set point. Enough cycles were tested so only small variances in energy 

consumption occurred between cycles, which ensured consistent idle performance and equal 

embodied tank energy at the start of each cycle. The energy use was summed from the start of 

the first burner cycle to just before the start of the final cycle and divided by the elapsed time to 

compute the average rate of energy use. Starting and ending the monitoring period just prior to 

the burner on time helped reduce the difference in the energy stored in the plant from the start to 

the finish of the monitoring period. The natural gas energy use was computed from the natural 

gas volumetric flow rate measured by a diaphragm flow meter multiplied by the heat factor (see 

equation 1). A watt transducer directly measured the electricity use.

To simulate the water conditions returning from a hydronic AH, the heating plant steady-state 

efficiency test was conducted using return water supplied at a steady temperature and flow rate.

The energy output rate was calculated from the water flow rate and temperature difference as 

shown by equation (2). A third physical plant test evaluated the transient performance of hot 

water delivery under three simulated conditions. The transient performance was quantified by 

determining the length of time required for the temperature to rise to within a specified 

difference from the set point or the maximum deviation from set point when a second load was

placed on the unit (e.g., a second shower starting while the first was in process). The 

measurement response time was minimized by using a resistance temperature detector in an 

emersion thermowell.

qin = Qg ∙ HF (1)

qout = Cp ∙ ρ ∙ Qw ∙ 60 ∙ ∆T (2)

η = qout/qin (3)

where:

qin = energy input, Btu/h

qout = energy output, Btu/h

Cp = specific heat of water (varies by temperature), Btu/(lb*°F)

ρ = water density at the flow meter (varies by temperature), lb/gal

Qw = water flow rate, gpm

∆T = temperature output difference, °F

Qg = burner natural gas flow rate, ft3/h

HF = natural gas heat factor, Btu/ft3

η = thermal efficiency
13

3.2 Air Handler Capacity Analysis

The hydronic AH tests characterized the performance of units in conditions within and outside 

the range of manufacturer specifications. Tests measured the energy output rate, supply air 

temperature, return water temperature, and electricity consumption for controlled variations in 

airflow rate, water flow rate, water supply temperature, and air inlet temperature. The energy 

output rate was computed from equation (2) using the water flow rate and temperature difference 

entering and leaving the hydronic coil.

3.3 Full System Analysis

Combi systems were run under several hourly load patterns to determine cyclical performance 

and maximum capacity. Several DHW and space heating load profiles were developed to 

evaluate the warmup and postpurge portions of the heating plant firing cycle. The tests required

the combi system to simultaneously supply heat for the equivalent of two showers and a space 

heat cycle. Equations (1) and (2) were used to compute the system energy input and output. The 

total output was computed from the sum of the space heating (hydronic AH) and DHW energy 

outputs. For many of the load patterns,

3 the system efficiency during cyclical operation was

computed from equation (3). The space heating energy output rate was computed from the water 

flow rate and temperature difference entering and leaving the hydronic coil and the DHW energy 

output calculation used the DHW flow rate and temperature difference between the water 

entering and leaving the heater.

3.4 Error Analysis

Two types of error must be accounted for in the laboratory measurements: accuracy of the 

sensors and error introduced by sensor location. Instrumentation was installed according to the 

manufacturers’ specifications such that additional error due to the installation was minimized or 

eliminated. Water temperature and flow sensors were installed with enough straight pipe length 

on the inlet and outlet of the sensor to maximize accuracy. Long straight ducts were not possible 

on the supply side of AHs. A three-sensor array was used to capture the temperature variance

across the supply duct. Accurately measuring air temperatures was difficult. Water temperatures 

and flow rates in the heating loop were therefore used to calculate the output delivered in space 

heating mode and water flow rates and temperatures in the hot water loop were used to calculate 

the output in hot water mode. Air temperatures were used for occupant comfort in the 

optimization process where being within a couple of degrees is acceptable.

The laboratory testing and optimization were mostly concerned with parameters that were 

measured directly, such as water temperature, water flow rate, and air temperature. The error in 

these parameters can be taken directly from the instrumentations (see Table 1. The error was 

calculated for three parameters: energy input (Qin), energy output (Qout), and efficiency (η).

Energy input had an error of 1.0%, accounting for the variation in the natural gas heating value 

and the error in the gas meter. Energy output had an error of 2.0% accounting for the error in the 

 3 The stored energy does not cause significant error in the efficiency estimate for patterns with higher loads. All test 

profiles were run from burner cut out to cut out, making the difference in tank temperatures small (assuming a 1°F 

temperature difference the stored energy change would be about 350 Btu). The effect of stored energy would be 

<1% for tests with an energy output >35,000 Btu.
14

water temperature difference, water flow rate, and the error in the estimation of the properties of

water. Combining these errors gives the efficiency an error of 2.2%.
15

4 Results

Laboratory tests were conducted between March and September 2011. All tests were conducted 

with ambient laboratory temperatures and return air temperatures of 68°F.

4.1 Standby Loss

Idle mode operation natural gas and electricity consumption data were collected at multiple 

temperature settings for all nine heating plants. Table 4 and Figure 7 show idle energy 

consumption at a set point of 140°F. Idle tests were run through enough burner cycles for 

consistency and repeatability. The actual test lengths varied by heating plant. Plants with larger 

storage capacities, storage WHs, were run longer, up to a week if necessary. Heating plants with 

no hot water storage (Boiler 1, Boiler 3, and TWH 1) had no gas use; the gas use for the units 

with storage varied from 29 to 103 therms/yr. The standby consumption of the three tank type 

WHs was about equal to or lower than that of TWH 2 and boilers with storage (Boiler 2 and 

Boiler 4). TWH 2, Boiler 2, and Boiler 4 had standby losses greater than desirable considering 

those units had notably smaller storage volumes than the tank type WHs (see Table 4). For these 

systems, the level of insulation appeared to have a greater effect than the storage size on idle 

losses. Figure 8, Figure 9, and Figure 10 show thermal images of heating plants Boiler 4, TWH 

2, and TWH 1, respectively. Boiler 4 had significant heat loss from the larger diameter,

uninsulated pipes inside the unit. The images of TWH 2 indicate that most of its heat loss was off 

the top of the internal tank. TWH 1 did not have storage, and therefore had small heat loss in 

standby mode. Reducing the temperature set point from 140°F to 120°F reduced natural gas 

consumption by 25% for heating plants with storage.

Table 4. Standby Site Energy Consumption at a Set Point Temperature of 140°F

Heating 

Plant

Storage

(gal)

Annual

Gas Use

Annual

Electricity Use

Total Annual 

Energy Use

(therms) ($) (kBtu) ($) (kBtu) ($)

Boiler 1 0 0 0 126.1 15 126 15

Boiler 2 6.6 36.6 37 141.0 17 3,796 53

Boiler 3 0 0 0 180.6 22 181 22

Boiler 4 14 102.5 103 153.4 18 10,408 121

THW 1 0 0 0 34.4 4 34 4

TWH 2 2 35.1 35 80.6 10 3,593 45

Tank 1 50 42.0 42 104.7 13 4,306 55

Tank 2 55 45.1 45 89.7 11 4,600 56

Tank 3 34 28.5 28 50.8 6 2,897 35

Note: Assumed rates of $1/therm and $0.12/kWh
16

Figure 7. Energy consumption during standby mode at a set point of 140°F

Figure 8. Infrared and visual images of Boiler 4 in standby operation

0

200

400

600

800

1000

1200

1400

Boiler 1 Boiler 2 

(156)

Boiler 3 Boiler 4 TWH 1 THW 2 Tank 1 Tank 2 Tank 3

Energy Consumption During Stand by [Btu/hr]

Gas Electric
17

Figure 9. Infrared and visual images of TWH 1 in standby operation

Figure 10. Infrared and visual images of TWH 2 in standby operation
18

4.2 Steady-State Efficiency

Inlet water temperature significantly impacts the efficiency of a condensing heating plant. This 

relationship has been previously demonstrated for condensing boilers (Arena 2011). The heating 

plant requires low return water temperature to operate in condensing mode. A series of 

laboratory tests determined the efficiencies associated when the hydronic AH return water 

temperature varied from 80°F to 120°F. Figure 11 illustrates the decrease in steady-state 

efficiency with increasing return water temperature. The efficiency of several heating plants 

decreased when the return temperature increased above 110°F. This can be significant over the 

life of the heating system. Assuming an annual load of 100 million Btu (a typical space heating 

load in Minnesota) a 10% efficiency reduction represents an increased gas cost of about $150/yr.

Figure 11. Heating plant steady-state efficiency over a range of return water temperatures

Figure 12 shows the steady-state efficiency4 of each heating plant for a range of return water 

temperatures (90°–120°F), which resulted in heating load of 20,000–100,000 Btu/h (a set point 

temperature of 130°F and a flow rate of 4 gpm were used for all tests). The heating plants had 

similar steady-state efficiencies with a low of 86.1% for Boiler 3 and high of 90.6% for TWH 2.

Given that the uncertainty of the efficiencies was 1.5 percentage points (see error bars in Figure 

12), the difference in steady-state efficiency was significant for the highest and lowest efficiency 

systems only. The average efficiency of the four boilers was 1.8 percentage points lower than the 

average for the TWHs and 1.9 percentage points lower than the tank type WH average. However, 

 4 Includes natural gas and electric energy input.

75%

80%

85%

90%

95%

100%

70 80 90 100 110 120 130

Steady-State Space Heating Efficiency

Return Water Temperature, °F

TANK 1 TANK 2 TANK 3 BOIL 1 BOIL 2 BOIL 3 BOIL 4 TWH 1 TWH 2

Set Temp = 130F, Flow = 4 Set Temp = 130°F,

Flow = 4 gpm
19

the difference in average efficiency between the three system types was within the expected 

measurement error.

Figure 12. Heating plant steady-state efficiency for 60 kBtu/h heating load

Steady-state efficiencies discussed in this section were those of the heating plants only and 

include natural gas and electricity energy consumption. They do not include any electricity

consumption that would be required by the AH to move the air.

Most heating plants tested used modulating burners, only TANK 1 and TANK 3 had single-stage 

burners (both at 100,000 Btu/h). Heating plants with modulating burners controlled their input 

rate to match the load. These burners reached a steady input and output rate. The steady 

performance was used for the test period. The two tanks with fixed burners cycled on and off to 

meet the load. Tanks were allowed to cycle until the total input and output of consecutive cycles 

were consistent.

All three tanks had maximum firing rates of 100,000 Btu/h and Boiler 1 had a maximum input 

rate of 75,000 Btu/h. These heating plants were unable to meet the 100,000 Btu/h load when 

return temperatures were reduced to 80°F.

4.3 Air Handler Capacity

Laboratory tests characterized AH performance and identified optimal space heating capacity 

ranges. The AH tests were configured to select the correct operational parameters for each AH.

The first criterion was that the systems produce a minimum supply air temperature of 110°F for 

an entering air temperature of 69°F, ensuring acceptable occupant comfort for the air delivered 

from supply resisters into the living spaces. The typical return air temperature in most Minnesota 
20 

homes was assumed to be 69°F; 110°F was the lowest delivered temperatures installers and 

program managers felt comfortable delivering to occupants. The minimum supply air 

temperature determined the lowest acceptable water flow rate for a given heating plant delivered 

water temperature and AH airflow rate. For example, Figure 13 shows that for AH7 the 

minimum allowable water flow rate was 2.2 gpm for an airflow rate of 1,050 cfm, 0.3 in. w.c.

static pressure across the unit, and a supply water temperature of 140°F. All the AHs installed in 

the laboratory had multiple motor speeds that could be selected manually, but were not 

continuously variable. The number of available airflow speeds varied from unit to unit. None of 

the AHs installed in the laboratory had the ability to vary air speed in response to the heating 

load.

5

Figure 13. AH7 performance mapping test of supply air temperature

To achieve high heating plant efficiency, a maximum return water temperature of 105°F was 

used to find the largest allowable water flow rate for each AH at a given airflow rate and 

supplied water temperature. The maximum flow rate allowable for AH7 at 1050 cfm was 2.6 

gpm (see Figure 14). The water flow rate range could be used with test data to determine the 

delivered capacity of each AH. AH7 had an optimized output capacity range of 45,000–50,000 

Btu/h at 1050 cfm (see Figure 15). Table 5 shows the acceptable flow ranges for each AH for a 

heating plant temperature setting of 130°F and several airflow rates. Tests were performed over a 

wide range of water flow and airflow rates and temperatures. Tests were then normalized for 

specific airflows to compare multiple units. Several of the AHs were not able to achieve 

performance that would allow for high efficiency heating plant operation (see Table 5). At a 

given airflow rate, an AH may not have been able to produce air temperatures higher than 110°F 

 

5 A unit that changes to a higher flow rate after 15 min of continuous operation was recently received and may be 

tested in the future.

80

90

100

110

120

130

140

0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5

Supply Air Temperature [F]

Flow Rate [GPM]

CFM=1050 Min Supply AirTemp

OK
21

and keep return water temperatures of 105°F. These tests were labeled not to have met 

performance parameters at that airflow. Some of the smaller AHs could not supply a large 

enough airflow to be tested at high flow rates. The equipment sizing process allowed an AH to 

be installed only if it was compatible with high efficiency heating plant performance for the 

home’s load.

Figure 14. AH7 performance mapping test of return water temperature

Figure 15. AH7 performance mapping test of space heating capacity

60

70

80

90

100

110

120

0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5

Return Water Temperature [F]

Flow Rate [GPM]

CFM=1050 Max Return Water Temp

OK

Airflow = 1050 CFM Max Return Water Temp

 

 

 

C 

0

10000

20000

30000

40000

50000

60000

0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5

AH output Capacity [Btu/hr]

Flow Rate [GPM]

CFM=1050

OK

Airflow = 1050 CFM

 

C 
22

Table 5. Sizing Chart for AHs for a 140°F Supply Water Temperature

AH

Airflow 

Rate

(cfm)

Water Flow Rate 

(gpm)

Heating Capacity 

(Btu/h)

Minimum Maximum Minimum Maximum

First Airflow

AH1 800 Never meets performance parameters

AH2 800 Never meets performance parameters

AH3 800 2.0 2.3 37,300 39,200

AH4 800 1.9 2.2 36,100 38,000

AH5 800 2.3 3.6 46,300 54,000

AH6 800 Never meets performance parameters

AH7 800 Was not tested at this flow rate

AH8 800 Never meets performance parameters

AH9 800 2.0 2.1 34,300 35,900

Second Airflow

AH1 1,100 Never meets performance parameters

AH2 1,100 2.3 2.8 46,800 48,000

AH3 1,100 2.0 2.8 46,600 47,500

AH4 1,100 Never meets performance parameters

AH5 1,100 2.0 3.2 53,000 56,300

AH6 1,100 Did not have enough airflow capacity

AH7 1,100 2.0 2.9 48,700 50,200

AH8 1,100 Did not have enough airflow capacity

AH9 1,100 Never meets performance parameters

Note: Data from several tests were normalized to common airflow speeds.

It was important to develop these curves for each AH. Installing improperly sized AHs would 

have produced an uncomfortable supply air temperature or high return water temperature. For 

example, a return water temperature of 120°F reduces the steady-state efficiency to <85% for all 

the heating plants and to almost 75% for the TWHs (see Figure 11). This reduced efficiency 

would increase annual gas costs $75–$150 for a typical house.

Figure 16 shows the AH6 laboratory test and manufacturer specification data for a supply water 

temperature of 140°F and 800 cfm airflow rate. The figure shows that the laboratory-measured 

temperature drop across the coil was within 6% of the manufacturers’ specifications. For all 

tests, the laboratory-measured performance was typically within 10% of the manufacturers’

specifications. This indicates that the manufacturer’s specifications could be used to estimate 

equipment performance. However, these specifications provide data for only a limited range of 

operation. The AHs had fixed-speed pumps, so data were typically provided for only a single 

water flow rate, which was always higher than what was expected to be used for the installed 

system. Extrapolating performance based on specification data would have increased the 

uncertainty of the range of acceptable flow rates. The manufacturers’ specifications often did not 

include air temperatures for the rated performance.
23

Figure 16. AH6 measured and manufacturer specified water temperature difference

4.4 Full System

Three DHW and space heating load profiles were developed for the full system tests to test a 

specific function of the combi system. Each was based on data collected from a previous WH

field study (Schoenbauer et al. 2010). The “Maximum Capacity” profile consisted of two 

showers and a space heating demand running simultaneously to test the high load capability of 

each system (Figure 17). The “DHW Interrupt” profile (Figure 18) was designed to access the 

impact of DHW temperature when a heating call interrupts a DHW event. Figure 19 describes 

the “Low Use” load profile that tested the system’s ability to provide hot water for short DHW 

draws.

Table 6 lists the heating capacity and duration of the events in the full system tests. 

0

10

20

30

40

50

60

0.0 0.5 1.0 1.5 2.0 2.5 3.0 3.5 4.0

Water temp drop across air handler [F]

Air handler flow rate [GPM]

Spec Sheet Lab Tests

 

Manufacturer Specifications

 

Results from Lab Tests
24

Figure 17. Maximum capacity test profile for full system tests

Figure 18. DHW interrupt load profile for full system tests

0

20

40

60

80

100

120

140

0 5 10 15 20

Demand [kBtu/hr]

Minutes

SPACE

DHW

0

10

20

30

40

50

60

70

0 5 10 15 20

Demand [kBtu/hr]

Minutes

SPACE

DHW
25

Figure 19. Low use load profile for full system tests

Table 6. Full System Test Load Profiles (Btu/h)

0

10

20

30

40

50

60

70

80

0 10 20 30 40

Demand [kBtu/hr]

Minutes

SPACE

DHW

TIME SPACE DHW SPACE DHW SPACE DHW

0:00:00 40000 0 0:00:00 0 0 0:00:00 0 0

0:02:59 40000 0 0:02:59 0 0 0:05:56 0 0

0:03:00 40000 70000 0:03:00 0 57500 0:05:57 0 70000

0:05:29 40000 70000 0:04:59 0 57500 0:06:37 0 70000

0:05:30 40000 125000 0:05:00 40000 57500 0:06:38 0 0

0:14:47 40000 125000 0:08:30 40000 57500 0:07:33 0 0

0:14:48 40000 55000 0:08:31 40000 0 0:07:34 0 37500

0:15:29 40000 55000 0:15:29 40000 0 0:07:54 0 37500

0:15:30 40000 0 0:15:30 40000 0 0:07:55 0 0

0:20:00 40000 0 0:19:59 40000 0 0:15:04 0 0

0:20:00 0 0 0:15:05 0 37500

0:15:17 0 37500

0:15:18 0 0

0:34:46 0 0

0:34:47 0 42500

0:35:41 0 42500

0:35:42 0 0

0:40:00 0 0

Total Btu 13,333 22,864 9,989 5,271 0 1,749

36,197 15,260 1,749

Max Capacity DHW Interrupt Low Use
26

4.4.1 Maximum Capacity

The “Maximum Capacity” profile consisted of two showers and a space heating load. The space 

heating load was consistent throughout the profile at 40,000 Btu/h.

6 The first shower was 2 gpm 

for 12 min (approximately 70,000 Btu/h). The second shower, starting 2.5 min after the first, was 

1.5 gpm and 10 min long (approximately 55,000 Btu/h). The systems were set to provide 120°F 

DHW supply water and the DHW inlet water temperature was about 60°F. Supply water 

temperatures from space heating and DHW outlets were compared at the beginning and end of 

the two-shower period. The beginning temperature was measured after the DHW outlet 

temperature stabilized for the first shower. Figure 20 shows the temperature comparison. The 

TWHs supplied the most consistent water temperatures, because the input ratings (199,000 

Btu/h) were significantly higher than the demand (165,000 Btu/h). The storage WHs showed a 

10°–30°F drop in DHW supply temperature and a 10°–30°F drop in space heating supply water 

temperature, which corresponded to a 5°–15°F drop in supply air temperature.

These systems showed temperature reductions when the storage capacity could no longer keep 

up with the difference in heating plant input (100,000 Btu/h) and desired load (165,000 Btu/h).

All four boilers had DHW priority controls. This control strategy diverts all heating capacity to 

the DHW side when there are simultaneous events. Boiler 2 and Boiler 3 were able to provide 

consistent DHW supply temperatures through the high demand period. Boiler 1 had a smaller 

input (75,000 Btu/h) than the DHW demand (125,000 Btu/h), and therefore could not meet the 

two-shower load. The internal heat transfer capabilities of Boiler 4 limited the maximum hot 

water flow rate at 2.9 gpm. Therefore, Boiler 4 could not meet the two-shower demand of 3.5 

gpm, which reduced the DHW supply temperature. These results were used to help develop the 

sizing recommendations presented in Section 4.6.

 6 A 40,000 Btu/h space heating load was representative of the design load requirements of the combi implementation 

project housing stock. All houses were required to go through the weatherization process. Weatherization provided 

improved air sealing, increased insulation, and provided other measures that lowered the homes’ space heating 

loads.
27

Figure 20. Difference in supply temperature from the start to end of a 

two shower and space heating demand

4.4.2 Response to Overlapping Events

The “DHW Interrupt” full system test profile was designed to test the effects of a space heating 

load initiated in the middle of a DHW draw. Changes in supply water temperature are especially 

important during showers. This draw pattern (Figure 18) consisted of 40,000 Btu/h space heating 

event starting 3 min into a 1.6-gpm shower event (17 min long at 57,500 Btu/h). Figure 21 shows 

the DHW temperature at the heating plant outlet for 10 s before and 15 s after the space heating 

load was added. The boilers and storage WHs all held constant outlet temperatures under the 

additional space heating load. The large quantity of stored hot water in the tank type WHs 

prevented a decrease in water temperatures. The boilers all have DHW priority, so when the 

space heating demand was initiated heat was not diverted from the DHW event. The DHW 

supply temperature for Boiler 4 was much higher than the set point (supply temperature/set point 

≈ 1.3) and is not displayed in Figure 21. This performance of Boiler 4 seems to be some kind of 

control issue, where water temperature is not reduced enough. An additional mixing valve would 

help produce a hot water temperature at the fixture that would be closer to the set point.

-70

-60

-50

-40

-30

-20

-10

0

10

TANK 1 TANK 2 TANK 3 TWH 1 TWH 2 BOIL 1 BOIL 2 BOIL 3 BOIL 4

Tamperature Change [F]

DHW Supply T SPACE Supply T Boilers have DHW priority Combined load ~20,000 Btu/hr for 10 min
28

Figure 21. DHW supply temperature response when space heating demand was initiated

There was a significant drop in the supply water temperature at the outlet tap of the two TWHs, 

but the drop was significantly lower after the mixing valve.

7 The maximum reduction was <5°F 

in both cases (TWH 1: 4.6°F and TWH 2: 3.6°F). Both TWHs had 2 min of inconsistent supply 

temperatures before returning to steady conditions. Figure 22 shows the TWH 1 supply water 

temperature at the outlet and after the mixing valve around the time of the space heating load.

There was a 5°F dip in the DHW temperature at the WH outlet, but that dip was reduced to about 

1°F after the mixing valve. The mixing valve also greatly reduced the length of the effect. The 

temperature became consistent at the mixing valve outlet in <30 s. The mixing valve had about 

15 ft of plumbing between it and TWH 1; this pipe length and the mixing value buffered the 

water temperature against the short-term transients.

 7 For the field monitoring portion of the project, all systems will be installed with mixing valves.

0.9

0.95

1

1.05

1.1

0 5 10 15 20 25

Supply Temperature/Set Point Temperature

Time [Seconds]

TANK 1 TANK 2 TANK 3

TWH 1 TWH 2 BOIL 1

BOIL 2 BOIL 3 BOIL 4

SPACE HEAT ON
29

Figure 22. DHW temperature changes before and after the mixing valve for TWH 1

4.4.3 Domestic Hot Water Supply Temperature

The “Low Use” load profile was used to evaluate each combi system’s ability to produce hot 

water for small DHW events. Table 7 describes each draw used in this profile. The total draw for 

each test was 0.75 gal. Hot water supply temperatures were measured within 6 in. of the heating 

plant outlet.

Table 7. DHW Draw Characteristics for the Low Use Load Profile

Time DHW Draw 

(gpm)

Load

(Btu/h)

5:57 2.0 70,000

6:37 Off 0

7:34 1.1 38,000

7:54 Off 0

15:05 1.1 38,000

15:17 Off 0

34:47 1.2 42,000

35:41 Off 0

Figure 23 shows the temperature profile at startup for Draw 4 with a 1.2 gpm flow rate. The tank 

type WHs were the fastest responding heating plants. These systems reached 95% of their set 

points in <5 s. Tank type WHs provide hot water at consistent temperatures at or below the set 

point, depending on the width of the dead band and the length of time since the burner last fired. 

110

115

120

125

130

135

140

145

0 5 10 15 20 25

Temperature [F]

Time [Seconds]

Temp after Mix Valve DHW Supply T SPACE Supply T SPACE HEAT ON
30

For tests of these tank type WHs, the dead band was small (3°–7°F) and the burners had recently

fired.

Figure 23. DHW supply temperature profile for a short 1.2-gpm draw

The two boilers with internal DHW tanks (Boilers 2 and 4) had temperatures higher than the set 

point for most draws. These heating plants store water at temperatures higher than the set point 

and mix in cold water to achieve the desired water temperature. Boiler 2 had a 10-s delay before 

it reached the desired temperature. Boiler 4 had a much longer delay, and the temperature was 

significantly over the set point for the duration of all four short draws of the “Low Use” profile. 

Boilers 1 and 3 were low mass boilers with no internal storage. These boilers had a ramp-up 

period before they reached the desired temperature.

With little or no storage capacity, increased delivery time is a concern for TWHs. The length of 

time a TWH requires to produce hot water relates directly to the length of time since the previous 

draw. If the heat exchanger is still filled with warm or hot water from a previous draw, hot water

delivery time will be shorter than if the water in the exchanger had cooled. For draw 4 of the 

“Low Use” profile, about 20 min had passed since the previous draw. TWH 1 produced hot 

water within 20 s. TWH 2 had 2 gal of hot water storage but took twice as long as TWH 1 to 

produce hot water on draw 4. Figure 24 shows the temperature profiles for all four draws in the 

“Low Use” profile. The figure illustrates that TWH 2 performed better in earlier draws in the 

profile. However, incoming cold water mixing with the stored hot water caused inconsistent 

initial output on all draws. Delivery times increased as the storage capacity was used.

-80

-60

-40

-20

0

20

40

60

0 5 10 15 20 25 30 35 40

DHW Supply Temp - Set Point Temp [F]

Time into Draw [Seconds]

TANK 1

TANK 2

TANK 3

TWH 1

TWH 2

BOIL 1

BOIL 2

BOIL 3

BOIL 4
31

Figure 24. DHW supply temperature profile for TWH 2

Table 8 summarizes the low use performance of each system during the draw period, including 

the ratio of supplied water temperature to the set point temperature, the volume of hot water 

delivered that was with 95% of the set point temperature, and the time it took for each draw to 

reach 95% of the set point temperature. The heating plants with no storage (Boiler 1, Boiler 3, 

and TWH 1) have longer delivery times, less volume delivered at temperature, and a lower 

average water delivery temperature than heating plants with storage.

Table 8. DHW Performance of Combi Systems Under Small and/or Short Draw Conditions

Heating 

Plant

Set 

Point

(°F)

Supply 

Temperature/

Space 

Heating

Volume 

Within 

95% of 

SP

Delay Until Within 95% of Set Point

(s)

1st

Draw

2nd

Draw

3rd

Draw

4th

Draw

Boiler 1 156 0.92 49% 33 4 >12 15

Boiler 2 120 1.03 96% 3 3 3 3

Boiler 3 120 0.84 27% 27 0 >12 3

Boiler 4 120 1.32 100% 0 0 0 0

THW 1 140 0.89 6% >40 >20 >12 19

TWH 2 140 0.92 47% 12 2 >12 41

Tank 1 140 1.00 97% 3 0 2 3

Tank 2 140 0.99 98% 5 1 4 4

Tank 3 140 1.01 97% 4 0 4 4

100

105

110

115

120

125

130

135

140

145

150

0 10 20 30 40 50 60 70

Outlet Temperature

Time Seconds

DRAW 1 DRAW 2 DRAW 3 DRAW 4

DRAW START
32

Note: Boilers 2, 3, and 4 have internal mixing valves that can be set by hand (Boiler 2) or by digital control (Boilers

3 and 4). Boiler 1 has no DHW temperature control; it is always set to 156°F and relies on an external mixing valve.

4.5 Installation Process and Specifications

The objective of the laboratory research was to develop guidelines for component selection, 

sizing requirements, and system installation. The results presented in previous sections were used 

to develop these guidelines and specifications. The first step in a combi installation is to measure 

or determine the key household characteristics: space and water heating loads, occupancy, 

distribution systems, and hot water end use.

The space heating load can be estimated from the building envelope characteristics and used to 

calculate the space heating design load. The home’s gas utility billing history and corresponding 

outdoor temperatures can help verify the space heating estimates. The water heating load can be 

estimated with a survey of end uses and number of residents. For sizing purposes, the large hot 

water draws are more important than the total DHW load. The shower events are typically the 

largest and have the greatest impact on occupant satisfaction. The actual shower hot water flow 

rates provide more reliable sizing estimates, but person-to-person variance of flow, cold water 

temperature into the home, and shower temperature selection necessitates the use of typical 

values. Low-flow showerheads were recommended to replace showerheads with flow rates >2.0 

gpm for the field phase of this project, to avoid excessive shower flow rates that cause occupant 

dissatisfaction.

Manufacturers of combi systems have not published sizing criteria. Contractors have 

traditionally sized systems on site. For this project the heating plant were sized to fit each home 

using the expected shower load and estimated space heating load. The natural gas burner input 

rate, storage capacity, and heating plant controls (DHW priority) are considered when sizing the 

system. Table 9 lists sizing guidelines for the three types of combi systems studied in this 

project. Each row corresponds to a heating plant that is being considered for use in the field 

monitoring phase of the project. The first four columns (type of heating plant, burner maximum

input, water storage capacity, and DHW priority) characterized the combi system. The next six 

columns describe six load profiles. These columns indicate whether the system will be properly 

sized for the specified loads. For example, a boiler with 199 kBtu/h input, zero storage, and 

DHW priority would have 150% of the capacity needed in a 40-kBtu/h, one-shower home 

(resulting in a 50% oversize factor). A heating plant was listed as “small” if the required output 

was greater than the available capacity, resulting in the system being undersized. The loads in 

Table 9 were selected to represent the range of the housing stock considered. When sizing was 

done for a specific home, the estimated loads of that home were used.
33

Table 9. Combi System Sizing Chart

Note: The SP value refers to the space heating load and 2 shw refers to two simultaneous showers.

An appropriate AH can be selected based on the space heating demand and heating plant. AH

testing is described in Section 4.3 and summarized in Figure 13, Figure 14, Figure 15, and Table 

5 can be used to select the correct AH. An AH with an optimized capacity range that matches the 

space heating demand of that home should be selected. The AH charts can then be used to select 

space heating water flow and airflow rates that allow for air temperatures >110°F with a return 

water temperature <105°F. These conditions will minimize the energy consumption of the combi 

Max InputStorage

kBtu/hr Gal 1 shw 2 shw 1 shw 2 shw 1 shw 2 shw

75 0 Yes small small small small small small

100 0 Yes 1% small 1% small 1% small

150 0 Yes 34% small 34% small 34% small

199 0 Yes 50% 13% 50% 13% 50% 13%

75 6 Yes small small small small small small

100 6 Yes 5% small 5% small 5% small

150 6 Yes 37% small 37% small 37% small

199 6 Yes 52% 15% 52% 15% 52% 15%

75 12 Yes small small small small small small

100 12 Yes 9% small 9% small 9% small

150 12 Yes 40% small 40% small 40% small

199 12 Yes 55% 17% 55% 17% 55% 17%

150 2 No 1% small small small small small

199 2 No 26% small 19% small 13% small

150 0 No 0% small small small small small

199 0 No 25% small 19% small 12% small

150 2 Yes 1% small small small small small

199 2 Yes 26% small 19% small 13% small

150 0 Yes 0% small small small small small

199 0 Yes 25% small 19% small 12% small

100 55 No small small small small small small

150 50 No 25% small 16% small 7% small

199 50 No 58% 5% 46% small 35% small

100 80 No 7% small small small small small

150 80 No 40% small 29% small 20% small

199 80 No 73% 15% 60% 9% 48% 4%

100 50 Yes small small small small small small

150 50 Yes 25% small 16% small 7% small

199 50 Yes 58% 5% 46% small 35% small

100 80 Yes 7% small small small small small

150 80 Yes 40% small 29% small 20% small

199 80 Yes 73% 15% 60% 9% 48% 4%

TANK

SP: 60,000 Btu/hr

Boiler

DHW 

priority?

SP: 40,000 Btu/hr SP: 50,000 Btu/hr

TWH
34

system. It may be necessary to allow for return water temperatures up to 110°F for some large 

demand applications. An increase in return water temperature would decrease system efficiency.

A DHW mixing or tempering valve will be included on every combi system in the field. A 

mixing valve blends the heating plant DHW supply water with cold inlet water to produce the 

hot water supplied to the fixtures. Space heating supply water temperatures are often required to 

be >120°F to meet the demand at an acceptable supply air temperature. A heating plant operating 

at temperatures >120°F carries an increased risk of scalding. Figure 23 and Figure 24 show 

supply water temperatures exceeding 140°F. The DHW loop of each system will include a 

mixing valve. The thermostatic nature of mixing valves also helped to reduce some of the 

transient changes in supply water temperature.

Several of the heating plants have outdoor temperature setback capabilities. This control method 

reduces the storage and supply water temperatures of the heating plant as the outdoor 

temperature increases. Figure 25 shows a typical outdoor temperature reset curve. Reducing the 

supply water temperature as the outdoor temperature increases helps match the space heating 

output to the house heating demand. Lowering the supply water temperature reduces the standby 

losses and increases the steady-state efficiency. The reduced space heating output would also 

increase the duration of the heating cycles. Outdoor reset has the potential to improve system 

efficiency in the shoulder and summer seasons. The 300 home implementation project will use 

resets where possible. Outdoor resets have two potential drawbacks that limit their use. Reducing 

the supply water temperature will reduce the supply air temperature and output capacity. Lower 

supply air temperatures could cause comfort problems, and less output capacity would increase 

setback recovery times. These two drawbacks must be considered in the design and optimization 

of the system if outdoor reset is to be used.

Figure 25. Typical outdoor setback temperature control curve

115

120

125

130

135

140

145

-20 0 20 40 60 80 100

Supply Water Temperature [F]

Outdoor Temperature [F]
35

5 Discussion

5.1 Air Handler Improvement

Most currently available hydronic AHs were not designed for combi systems with a condensing 

heating plant. Therefore, low return water temperatures were not prioritized and there was little 

cross-flow performance of the coils. Counter-flow heat exchangers are designed with the water 

flow and airflow paths in opposite directions, so that the water leaving the heat exchanger 

transfers heat to the entering cool air. The hydronic AHs were configured with counter-flow 

patterns, but the coil geometries did not produce counter-flow performance. Only one AH was 

able to deliver air temperatures more than a couple degrees above the return water temperature.

This was significant because the conditions required for optimized heating plant efficiency and 

homeowner comfort (in retrofit applications) were <105°F return water temperature and >110°F 

delivered air temperature.

All potable rated hydronic AHs had the heating loop pump inside. Building codes require an 

integral control to prevent coil water stagnation by running the circulating pump when there is a

long period with no heating demand. All but one of these pumps had a single speed that could 

not be adjusted. Fixed-speed pumps require a balancing valve in the circulation loop to reduce 

the water flow rate for optimal AH and heating plant performance. The only variable-speed

pump was manually controlled, forcing the installer to select the speed. This was an 

improvement over the fixed-speed pumps, but could have been even better had the pump

controls changed the pump speed with demand so an increased output could be delivered on 

colder days or during periods with higher demand (e.g., recovery from thermostat setback).

All the AHs had multiple fan speeds. The range of available speeds often increased as coil size 

increased, creating a situation where it was difficult to deliver acceptable air temperatures in the 

high capacity units. Once the fan speed was selected, it did not adjust to load changes. A simple 

control could allow an AH to meet a large load (e.g., morning setback) more quickly. This 

control could increase the fan speed once a certain runtime was met. For example, if the base fan

speed produced 800 cfm and there was a continuous call for heat for 10 min, the fan speed would 

increase to deliver 1200 cfm.

5.2 Heating Plant Improvement

The primary/secondary boiler loop configuration presents a significant obstacle to high 

efficiency space heating. Figure 26 shows the combi boiler installation from one manufacturer.

All combi boiler manufacturers have similar plumbing diagrams. Manufacturers require the 

primary/secondary loop to control the flow rate of water through the boiler. There is a concern 

that overheating could damage the boiler, even causing it to explode, if the flow rate is too low 

and air is in the water loop. This configuration increased return water temperatures. There was 

mixing within the primary loop while meeting a space heating demand. Flow rates in the primary 

loop, typically a fixed rate set by the boiler, and the secondary loop, a fixed rate set by the AH, 

affected the extent of mixing. The system efficiency optimization typically specified flow rates 

in the secondary loop to be about half the flow rate of the primary loop. This setup allowed for a 

large amount of mixing and an estimated 5°–20°F increase between the water temperature 

returning from the AH and entering the boiler. The magnitude of the temperature change 

depended on the flow rates, water temperatures, and plumbing configuration. This temperature 
36

increase related to about a 3%–11% system efficiency reduction, or a potential $40–$200 annual 

increase in space heating energy cost (assuming $1/therm of natural gas).

Figure 26. Typical combi system installation with a boiler
37

6 Conclusions

The highest system efficiencies were achieved by minimizing the water temperatures returning 

from the hydronic AH. Heating loads were determined for each site. These loads were used to 

select the water flow rates and airflow rates that would result in low return water temperatures 

while meeting the necessary load with an acceptable air temperature. Laboratory tests showed 

that the heating plant steady-state efficiency decreased with increasing return water temperature.

The decrease in efficiency became more significant as the return temperature increased above 

110°F.

Laboratory testing verified that heating loads up to 50,000 Btu/h with acceptable return water 

temperatures and supply air temperatures. These designs provide steady-state space heating 

efficiencies >85%.

System design and sizing information developed in the laboratory is currently being used to 

optimize systems for the 300 site implementation project. The laboratory work has been 

invaluable to this process. The installers and program managers have confidence that these 

systems will meet the needs of the homeowners because of laboratory tests.

The performance of combi systems is limited by the currently available equipment. Several 

manufacturers, at least one in direct reaction to the findings from this project, have begun to 

improve combi equipment. AHs with improved heat transfer performance will allow for lower 

and lower return water temperatures. Improved controls will enable adjustment and continuous 

optimization.
38

7 References

Arena, L. (2011). Building America Systems Evaluations: In‐Field Performance of Condensing 

Boilers in Cold Climate Region. Steven Winters Associates, Inc.

Bohac, D.; Hancock, M.; Dunsworth, T.; Makredes, S. (1995). Energy Savings and Field 

Experiences with Dual Integrated Appliances. Proceeedings paper. Minneapolis, MN: 

Center for Energy and Environment, August 1. http://aceee.org/proceedingspaper/ss94/panel03/paper02.

Butcher, T. (2011). “The DHW Performance of Heating Boilers.” Presented at the ACEEE Hot 

Water Forum, May 11, Berkeley, CA. 

www.aceee.org/files/pdf/conferences/hwf/2011/7A%20-%20Tom%20Butcher.pdf.

Energy Conservatory. (2008). TrueFlow Air Handler Flow Meter. 

www.energyconservatory.com/download/trueflow.pdf.

Rudd, A. (2010). Building America Special Research Project: Space Conditioning Systems for 

High Performance Home. Research report. Building Science Corporation, November 15.

Schoenbauer, B, Hewett, M.; Bohac, D. (2011). “Actual Savings and Performance of Natural 

Gas Instantaneous Water Heaters.” ASHRAE Transactions 117 (1): 657–672.

Thomas, M. (2011). “Designing a Highly Efficient Combination Space/Water Heating System.”

Presented at the ASHRAE Annual Conference, June 26, Montreal.
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