residential ventilation: a review of established … · 4.3.1 test set up and methodology ..... 22...
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Residential Ventilation: A Review of Established Systems and a Laboratory Investigation of the Fine Wire Heat
Recovery Ventilator
by
Samuel van Berkel
A thesis submitted in conformity with the requirements for the degree of Master of Applied Science
Department of Civil Engineering University of Toronto
© Copyright by Samuel van Berkel 2014
ii
Residential Ventilation: A Review of Established Systems and a
Laboratory Investigation of the Fine Wire Heat Recovery
Ventilator
Samuel van Berkel
Master of Applied Science
Department of Civil Engineering
University of Toronto
2014
Abstract
The fine wire HRV is a novel concept for decentralized residential ventilation heat recovery
using thin copper wires to transfer sensible energy between supply and exhaust airstreams. The
HRV can be incorporated into the building envelope, in effect creating a “breathing wall” ideally
suited to demand controlled ventilation.
Performance testing conducted in the laboratory indicates that fan electricity consumption was as
low as 1.1 W per L/s, while sensible heat recovery efficiency was as high as 82%. Overall, the
fine wire HRV is comparable to the top 5% of HRVs available in North America. When used in
conjunction with demand controlled ventilation, the modeled ventilation heating load was
reduced by 61% and the total heating load was reduced by 17%. Fan electricity consumption was
also reduced by 61%, corresponding to a reduction in household electricity use of roughly 5%.
Additional modeling and field installations are recommended.
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Acknowledgments
I would like to thank Professor K.D. Pressnail and Professor J. Siegel for their ideas, support and
guidance in writing this thesis. I would also like to thank Doug Hart for encouraging me to
continue my studies and first bringing the fine wire heat exchanger to my attention. Finally I
would like to thank Eur van Andel, inventor of the fine wire heat exchanger, for supplying a heat
exchanger sample and for his guidance constructing a prototype heat recovery ventilator.
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Table of Contents
1 Introduction ................................................................................................................................ 1
2 Review of Existing Technologies .............................................................................................. 3
2.1 Distribution System ............................................................................................................ 3
2.1.1 Centralized Ventilation ........................................................................................... 4
2.1.2 Decentralized Ventilation ....................................................................................... 5
2.2 Ventilation Control ............................................................................................................. 6
2.3 Energy Recovery Technology ............................................................................................. 8
2.3.1 Plate ......................................................................................................................... 9
2.3.2 Rotary .................................................................................................................... 10
2.3.3 Run-Around .......................................................................................................... 10
2.3.4 Heat Pipe ............................................................................................................... 11
2.3.5 Alternating Flow Regenerators ............................................................................. 11
3 The Fine Wire HRV ................................................................................................................. 12
3.1 Heat Recovery Technology ............................................................................................... 12
3.2 Enclosure and Manifold Design ........................................................................................ 13
3.3 Distribution System .......................................................................................................... 15
3.4 Ventilation Control ........................................................................................................... 16
4 Performance Measurements ..................................................................................................... 17
4.1 Pressure Drop .................................................................................................................... 17
4.2 Fan Electricity Consumption ............................................................................................ 19
4.3 Airstream Cross-Leakage .................................................................................................. 22
4.3.1 Test Set Up and Methodology .............................................................................. 22
4.3.2 Measurements ....................................................................................................... 23
4.3.3 Estimating Cross-Leakage .................................................................................... 25
4.3.4 Limitations ............................................................................................................ 27
4.4 Sensible Heat Recovery Efficiency .................................................................................. 27
4.4.1 Measurement Standards ........................................................................................ 27
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4.4.2 Test Set Up and Methodology .............................................................................. 29
4.4.3 Results ................................................................................................................... 31
5 Preliminary CFD Model ........................................................................................................... 34
6 Whole-Building Model ............................................................................................................ 36
6.1 Indoor Air Quality Metrics ............................................................................................... 36
6.1.1 Fine Particulate Matter (PM2.5) ............................................................................. 36
6.1.2 Formaldehyde (HCHO) ........................................................................................ 37
6.1.3 Ozone (O3) ............................................................................................................ 37
6.2 Methodology: Energy Use ................................................................................................ 38
6.3 Methodology: Indoor Air Quality ..................................................................................... 39
6.3.1 Fine Particulate Matter (PM2.5) ............................................................................. 41
6.3.2 Formaldehyde (HCHO) ........................................................................................ 41
6.3.3 Ozone (O3) ............................................................................................................ 42
6.4 Analysis ............................................................................................................................. 43
6.4.1 Constant Ventilation Rate ..................................................................................... 43
6.4.2 Variable Ventilation Strategy ............................................................................... 43
6.4.3 Occupancy Schedules ........................................................................................... 46
6.5 Results ............................................................................................................................... 48
6.5.1 Energy Consumption ............................................................................................ 48
6.5.2 Indoor Air Quality ................................................................................................. 50
6.6 Limitations ........................................................................................................................ 51
7 Discussion ................................................................................................................................ 52
8 Conclusions .............................................................................................................................. 54
9 Future Research ........................................................................................................................ 55
9.1 Condensation and Freezing ............................................................................................... 55
9.2 Field Installation ............................................................................................................... 56
Appendix A: Photos of Prototype, Testing and Installations ....................................................... 62
Appendix B: Sensible Heat Recovery Efficiency Test Data ....................................................... 66
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List of Tables
Table 1. Mass balance model parameter values ............................................................................ 40
Table 2. Pollutant concentrations at constant ventilation rates ..................................................... 43
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List of Figures
Figure 1. Common residential ventilation systems ......................................................................... 1
Figure 2. Conventional heat/energy recovery ventilator ................................................................. 2
Figure 3. Types of heat/energy recovery ventilators ...................................................................... 9
Figure 4. Fine wire heat exchanger and manifold ......................................................................... 12
Figure 5. Heat exchanger and heat recovery process .................................................................... 13
Figure 6. Photos of variable speed brushless DC centrifugal fans and air sealing ....................... 14
Figure 7. HRV enclosure and heat exchanger manifold ............................................................... 14
Figure 8. Typical use of the fine wire HRV in a bedroom ........................................................... 15
Figure 9. Fine wire HRV concept and distribution system ........................................................... 16
Figure 10. Test setup for pressure and flow rate measurements ................................................... 18
Figure 11. Pressure and flow rate measurements used to develop system curve ......................... 19
Figure 12. Fine wire HRV electrical power consumption at different flow rates ......................... 20
Figure 13. Fine wire HRV ventilation electrical efficiency at different flow rates ...................... 21
Figure 14. Distribution of electrical power consumption of HRVs .............................................. 21
Figure 15. Dry ice and configuration of leakage testing using CO2 ............................................. 23
Figure 16. CO2 concentrations in Airstreams A and B ................................................................. 24
Figure 17. CO2 concentrations in Airstream B before and after heat exchanger .......................... 24
Figure 18. Process for estimating CO2 concentrations in Airstream A ........................................ 26
Figure 19. Location of temperature measurements in standard test procedures ........................... 29
Figure 20. Heat recovery efficiency test setup and location of thermocouples ............................ 30
Figure 21. Temperature measurements at two airflow rates ......................................................... 31
Figure 22. Effect of heat loss through the HRV enclosure ........................................................... 32
Figure 23. Sensible heat recovery efficiency test results .............................................................. 33
Figure 24. Distribution of sensible heat recovery efficiencies of HRVs ...................................... 34
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Figure 25. Initial CFD modeling results ....................................................................................... 35
Figure 26. Variable ventilation strategy ........................................................................................ 44
Figure 27. Required duration of increased mechanical ventilation .............................................. 45
Figure 28. Occupancy schedules ................................................................................................... 46
Figure 29. Living room pollutant concentrations resulting from occupancy schedules ............... 47
Figure 30. Bedroom pollutant concentrations resulting from occupancy schedules .................... 48
Figure 31. Energy use of constant and variable ventilation strategies .......................................... 49
Figure 32. Pollutant concentrations of constant and variable ventilation strategies ..................... 50
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1 Introduction
The energy efficiency of buildings has steadily increased since the energy crisis of the 1970’s
prompted improvements in construction methods and HVAC technology (Wray et al. 2000). A
major component of this improvement in energy efficiency has come as a result of reduced
uncontrolled air leakage through the building envelope. Over the same period, changes in
building materials, appliances, home furnishings and manufactured products have resulted in
new types of indoor pollutants and increased emissions levels (Sherman and Walker 2007). As a
result, operable windows and air leakage can no longer be relied upon to provide adequate
residential ventilation, particularly in cold climates during the winter.
In order to provide a healthy indoor environment for building occupants, most jurisdictions
prescribe residential ventilation rates based on the size of the space and the number of
anticipated occupants. These ventilation rates are intended “to provide indoor air quality that is
acceptable to human occupants and that minimizes adverse health effects” (ANSI/ASHRAE
2013). Builders have traditionally met the requirements with central exhaust-only systems, which
are relatively simple to install and have a low initial cost (Wray et al. 2000). As interest in energy
conservation grows, balanced supply and exhaust systems are becoming increasingly popular in
cold climates because they allow for heat to be recaptured from the exhaust air. Without
ventilation heat recovery, energy savings from improvements in the air-tightness of the building
envelope are offset by heat losses due to increased mechanical ventilation. Balanced ventilation
systems also allow for pre-filtration of supply air and prevent depressurization, which can have
negative effects on indoor air quality (Russell et al. 2005).
(a) Exhaust-only (b) Balanced supply & exhaust
Figure 1. Common residential ventilation systems (Oikos Green Building Source 1995)
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Balanced ventilation systems recover energy from the exhaust air using a heat recovery
ventilator (HRV) or energy recovery ventilator (ERV). The difference between HRVs and ERVs
is described in Section 2.3. Both are packaged ventilation units which provide both supply and
exhaust airflows while using a heat exchanger to transfer energy between the two airstreams, as
illustrated in Figure 2. In cold outdoor conditions, the energy captured from the outgoing exhaust
airstream pre-heats the incoming supply airstream, reducing the additional space heating load
created by the ventilation air.
Figure 2. Conventional heat/energy recovery ventilator
HRVs and ERVs provide a solution for maintaining indoor air quality while achieving the energy
benefits of an increasingly airtight building envelope. However, they have also introduced a new
set of challenges. Additional electricity is required to power fans, and in some cases this
electricity consumption can partially or wholly offset the energy benefits of reduced space
heating and cooling, particularly if the system is run continuously (El Fouih et al. 2012). Even in
the winter, when fan electricity can partially offsets space heating requirements, the high cost of
electricity relative to natural gas results in increased energy costs. Energy recovery efficiency
also varies widely between commercially available units, with many HRVs recovering less than
60% of the available heat (Home Ventilation Institute 2014). In houses without forced air heating
systems, additional ductwork is required, utilizing space and creating interconnections which
allow noise to travel between rooms. HRV and ERV controls can also be confusing for home
owners, leading to over- or under-ventilation in some circumstances.
While these challenges are being partially addressed by performance improvements of
conventional HRVs and ERVs, further progress may require completely reimagining residential
ventilation and energy recovery. This thesis was motivated by one such effort in the Netherlands.
Indoor Air
(warm & stale)
Pre-Heated
Supply Air
(warm & fresh)
Outdoor Air
(cold & fresh)
Exhaust Air
(cold & stale)
Heat Exchanger Supply Fan
Exhaust Fan
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In the 1990s a Dutch company called Fiwihex developed an HRV which uses thin copper wires
to transfer sensible energy between supply and exhaust airstreams and termed their invention a
fine wire heat exchanger (2006). Because of the efficient heat transfer mechanism, the heat
exchanger profile can be constructed thin enough to allow the HRV to be incorporated into the
building envelope, in effect creating a “breathing wall” ventilation system without ductwork.
In this initial investigation of the fine wire HRV, its efficiency and performance are measured in
the laboratory and its potential application for decentralized ventilation heat recovery in
residential buildings is evaluated. Other ventilation technologies and strategies are also reviewed
to provide the reader with perspective and a basis for comparison.
2 Review of Existing Technologies
Ventilation and air leakage are estimated to account for a third of all energy used for space
conditioning (Liddament and Orme 1998). In North America, ventilation rates are established by
ANSI/ASHRAE Standard 62.1-2013 – Ventilation for Acceptable Indoor Air Quality and
ANSI/ASHRAE Standard 62.2-2013 – Ventilation and Acceptable Indoor Air Quality in Low-
Rise Residential Buildings. Although these standards are well established, there remains
significant variation in how much energy ventilation systems use and the quality of indoor air
they provide. Sherman and Walker (2007) compared common residential ventilation systems
meeting ANSI/ASHRAE Standard 62.2 and found differences of two or three times in total
ventilation energy use, suggesting that there are still significant improvements that can be made
to residential ventilation design, both by optimizing existing systems and by exploring new
concepts in residential ventilation. This section provides a review of existing residential
ventilation technologies by focusing on three key areas: distribution system, ventilation control,
and energy recovery technology.
2.1 Distribution System
Russell et al. (2005) provide a thorough review of existing residential ventilation strategies.
Mechanical ventilation may be exhaust-only, supply-only or balanced supply and exhaust, and
may be operated either continuously or intermittently. Natural ventilation strategies include
operable windows, passive stack ventilation and solar chimneys. Limited energy recovery is
possible in naturally ventilated systems while a heat pump can be used to recapture energy from
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exhaust-only mechanical systems. However, for most practical applications ventilation heat
recovery is limited to balanced supply and exhaust mechanical systems (Russell et al. 2005).
Balanced supply and exhaust systems may be centralized, having a single set of supply and
exhaust fans for the whole single family detached house. In the case of multi-unit residential
buildings, multiple centralized supply and exhaust fans are used and each serves a number of
suites. Decentralized ventilation makes use of individual supply and exhaust fans to ventilate
individual rooms or zones of the house, or in the case of multi-unit residential buildings,
individual supply and exhaust fans for each suite.
2.1.1 Centralized Ventilation
In multi-unit residential buildings, the most common balanced ventilation strategy in older
buildings is to have central air handlers supply air to the corridors and exhaust air from
bathrooms and kitchens. Although some portion of the supply air is expected to enter suites
through the pressurized corridors, these systems are mostly designed to prevent the movement of
contaminants into the corridors and between suites. Infiltration through walls and operable
windows is required to supply much of the fresh air for these buildings, often leading to
inadequate indoor air quality and comfort. In newer multi-unit residential buildings, fresh air is
typically delivered directly to suites in addition to corridors, improving indoor air quality but
increasing cost and space required for ductwork. In either case, centralized systems rely on
interconnections between suites, making it more difficult to control pressure differences induced
by stack effect, as well as the spread of odours, sound, pests, fire and smoke (CMHC 2003).
In single family homes, balanced supply and exhaust systems with ventilation heat recovery can
be installed in a number of ways, each having advantages and disadvantages as described by
Building Science Corporation (2013). In a single-point system, air is typically delivered to a
single room, and exhausted from another room. This system requires the least ductwork and the
simplest controls, making it the most economical. It does not distribute air throughout the house
or provide mixing however, important factors in achieving adequate indoor air quality in all
areas of the home (Rudd and Lstiburek 2000). A multi-point system typically delivers air to each
bedroom and living room, and exhausts air from bathrooms and kitchens. Although the multi-
point system achieves good air distribution, it has a high installation cost because of the
extensive ductwork required. In houses with a central air handler, fresh air can be delivered to
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the air handler’s supply trunk, and exhaust air can be extracted from return trunk. This integrated
system provides the distribution benefits of a fully ducted system, without the additional
ductwork costs. An integrated system is more vulnerable to control issues since the ventilation
control must be interconnected with the air handler control (Building Science Corporation 2013).
2.1.2 Decentralized Ventilation
Decentralized ventilation has a number of advantages over centralized ventilation. From the
perspective of indoor air quality, decentralization allows a ventilation system to meet the specific
demand of each zone without having to balance different requirements throughout the building.
Such systems can respond to increased occupant-related pollutant emission rates by increasing
ventilation rates when occupancy is high and can conserve energy by reducing ventilation rates
when the zone is unoccupied. By providing increased ventilation near pollutant sources,
decentralized systems can also achieve higher ventilation effectiveness than mixed ventilation
systems, reducing the volume of supply air required and the energy required to condition and
deliver that air. By compartmentalizing buildings into zones, decentralized ventilation also
allows for increased temperature variation in unoccupied spaces, further reducing space
conditioning requirements (Knoll 1992). In multi-unit buildings, compartmentalizing individual
suites reduces horizontal and vertical air movement, reducing air leakage through the building
envelope, improving comfort, and reducing space conditioning requirements (CMHC 2005).
By decentralizing ventilation systems, building occupants are able to have greater control over
their indoor environment, improving comfort and overall satisfaction. Elimination of a central air
distribution network also reduces space requirements for ductwork. It can mean reduced sound
transmission between rooms or suites (Knoll 1992), and in the case of multi-unit buildings, better
control over the spread of pests, fire, and smoke (CMHC 2003).
Since individual ventilation systems are required for each room or suite, decentralized ventilation
is typically more expensive to install and to maintain than a centralized system. Further,
depending on the location of ventilation units and the types of fans used, occupants may be
exposed to greater noise levels.
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2.2 Ventilation Control
Residential ventilation requirements calculated using ANSI/ASHRAE Standard 62.1 and 62.2
include minimum constant whole-building ventilation rates based on the floor area and the
number of intended residents. The simplicity of this constant ventilation approach makes it
relatively straight forward for builders to understand and implement, and ensures adequate
indoor air quality is provided under most circumstances.
At the same time, because of the significant energy cost associated with the delivery and
conditioning of ventilation air, there has long been an interest in finding strategies to reduce
overall ventilation rates while maintaining adequate indoor air quality. Because pollutant
emission rates vary, and spaces are not continuously occupied, constant ventilation inevitably
means ventilating more than necessary at some times. As early as 1990, the need for demand-
controlled ventilation (DCV) had been identified (International Energy Agency 1990), as it was
clear that by continuously adjusting the fresh air delivery to meet the demand, energy
consumption could be reduced while indoor air quality could be maintained or even improved.
In some cases, DCV can be achieved by increasing or decreasing ventilation based on a simple
schedule. Scheduled ventilation is appropriate in buildings with predictable occupancy patterns,
such as retail spaces or office buildings which are unoccupied at night or during the weekends,
and is uncommon in residences. Scheduled ventilation is the lowest cost DCV strategy, but is
limited by how much energy it can save because it must make fairly conservative assumptions
about how long and to what fraction spaces are occupied. Scheduled ventilation may also fail to
provide adequate indoor air quality if the programming does not consider atypical occupants
such as nighttime cleaners, or if there is an unexpected deviation from the assigned schedule.
Sensor-based demand-controlled ventilation (SBDCV) provides a more adaptable option than
simple ventilation scheduling. Fisk and De Almeida (1998) identified three characteristics of
applications best suited to air quality SBDCV: (1) a pollutant or ‘indicator’ exists which is
relatively easy to measure and which dominates such that reducing its concentration with
ventilation will provide sufficient control of all other pollutants; (2) large spaces where pollutant
emission rates or occupancy are unpredictable and vary with time; and (3) locations with high
heating or cooling loads and high energy costs.
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The most common parameters used for SBDCV is carbon dioxide (CO2), which is continuously
exhaled by humans and is relatively inexpensive to monitor. While recent research by (Satish et
al. 2012) suggests CO2 may impair cognitive performance at concentrations as low as 1000 ppm,
it has traditionally been used as an indicator of bioeffluents and other emissions related to
occupant activity, rather than a pollutant itself (Fisk and De Almeida). The ventilation rate per
person roughly correlates to the steady state CO2 concentration in a space, and this principle is
used to estimate the number of occupants and set the ventilation rate accordingly. Since CO2
sensors have some lag time, and ventilation rates and occupancy are highly dynamic processes, a
basic set-point control strategy is often ineffective, and a proportional or exponential control
strategy is preferable (Schell et al. 1998). Recent updates to ANSI/ASHRAE Standard 62.1 only
permit a basic CO2 set-point control strategy when used very conservatively, resulting in
significant over-ventilation (Dougan and Damiano 2011).
Dougan and Damiano identify several challenges in implementing CO2-based SBDCV, most
significantly sensor precision (±50 ppm or greater sampling error), drift and time lag, but also
ventilation airflow rate accuracy and changes in outdoor CO2 concentration. They note that this
can lead to an accumulated root-mean-square error of up to 50%. As a result, many designers act
conservatively, leading to over-ventilation and increased energy consumption, often wholly or
partially defeating the benefits of DCV. CO2-based SBDCV is also unable to respond to
pollutant emissions which are not correlated with occupancy, such as emissions from building
materials and furniture (Fisk and De Almeida 1998).
Because of the limitations of CO2-based SBDCV, a number of alternative strategies for detecting
occupancy and controlling occupant-generated pollutants have been developed. Naghiyev et al.
(2014) compared three unobtrusive occupant detection techniques in a residential setting include
CO2, passive infra-red (PIR) and device-free localization (DfL). PIR technology functions by
sensing human body temperature against the temperature of background objects. The main
advantage of PIR sensors is that they are low cost, easy to use and offer immediate feedback.
The main disadvantage is that they cannot be used to estimate the number of occupants. DfL
functions by detecting the absorption of radio signals by the human body. It is possible, although
difficult, to estimate the density of occupants using DfL.
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The study by Naghiyev et al. (2014) found that all three occupancy detection techniques
functioned to some degree. The CO2 sensors suffered from slow response times, while the PIR
sensors failed to detect reduced-movement occupants in some circumstances. DfL technology is
promising, but further research is required to determine how best to configure the sensors and
analyze the data. CO2 and PIR sensors can be used in combination, although both are influenced
by the metabolic rate of occupants.
In residential settings, humidity-based DCV can be effective for controlling moisture. It is most
often implemented in bathrooms or kitchens, and used to activate exhaust fans. Humidity-based
SBDCV must be combined with another ventilation strategy, as humidity is not a good indicator
for most residential pollutants (Fisk and De Almeida 1998).
Total volatile organic compound (TVOC) sensors can offer more precise control of non-
occupant-based emissions, such as those from building materials and furniture. Unfortunately,
controlling ventilation based on TVOCs is problematic because of the large variability in toxicity
of individual VOCs relative to each other (Fisk and De Almeida). In residential settings,
emissions from building materials and furniture are typically stable enough that TVOC-based
SBDCV has limited benefit.
2.3 Energy Recovery Technology
As described by Kakaç and Liu (2002), heat exchangers can be classified into two basic
categories based on the process of heat transfer: recuperative and regenerative. Recuperative heat
exchangers are characterized by one fluid continuously recovering heat directly from the other
fluid, either by direct contact between the fluids or, in the case of ventilation heat recovery,
through a separating wall. Regenerative heat exchangers use an intermediate storage medium to
transfer energy to and from the two airstreams. The airstreams may alternately occupy a single
flow passage containing the storage medium or the storage medium may move between the two
airstreams.
Heat exchangers used for ventilation heat recovery can be further classified into those which
recover sensible energy only (referred to as “heat recovery ventilators” or HRVs) and those
which recover sensible energy and latent energy (referred to as “energy recovery ventilators” or
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ERVs). ERVs are most commonly used in hot climates where the latent portion of the cooling
load dominates, while HRVs are more common in cold climates where heating load dominates.
Plate, rotary and run-around heat exchangers are the conventional technologies most commonly
used to recover energy from ventilation air. Of the three conventional technologies, plate heat
exchangers are the most common technology, due largely to their simplicity and relatively low
cost. Rotary heat exchangers are most common where space is available and significant latent
energy recovery is desired. Run-around systems have lower efficiencies, but allow for heat
recovery between physically separated air streams.
Heat pipes and alternating flow regenerators are examples of more experimental technologies
which may be appropriate for specific applications, but have not been widely adopted in North
America. While alternating flow regenerators are promising for decentralized ventilation, there
has been limited peer-reviewed study of the technology.
Plate Source: www.innergytech.com
Rotary Source: www.innergytech.com
Run-Around Source: www.immak.eu
Heat Pipe Source: www.immak.eu
Alternating Flow
Regenerator Source: www.lunos.de
Figure 3. Types of heat/energy recovery ventilators
2.3.1 Plate
The most common form of HRV for buildings is a recuperative plate heat exchanger, sometimes
referred to as a cross-flow heat exchanger because the two airstreams are brought together in a
cross or counter-flow arrangement. Plate heat exchangers have the advantage of being relatively
compact, low-cost and free of moving parts. Plate heat exchangers designed to recover sensible
energy only are typically constructed from aluminum or polypropylene. Although aluminum has
the advantage of higher thermal conductivity, polypropylene cores weigh less, are more
corrosion and chemical resistant, and collect fewer fouling particles (Mardiana and Riffat 2013).
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Flat plate ERVs designed to recover both sensible and latent energy have been built from
vapour-permeable paper sheets for over 30 years (Osamu 1984). Zhang and Jiang (1999) tested
membrane-based flat plate heat exchangers and concluded that, under hot and humid conditions,
they could recover twice as much latent energy as the existing paper systems. Both technologies
allow the movement of water vapour between air streams, while preventing the transfer of
pollutants. Today both vapour-permeable paper and membrane plate technologies are in use, but
are less common than rotary heat exchangers when moisture transfer is desired.
2.3.2 Rotary
Rotary heat exchangers, often referred to as heat wheels or energy wheels, are the most common
form of regenerative heat exchangers used for ventilation heat recovery. The matrix or storage
medium is in constant movement, passing periodically from the hot airstream to the cold
airstream. Early rotary heat exchangers used steel wool to store sensible energy, while newer
models make use of silica-gel desiccants to store both sensible and latent energy (Nóbrega and
Brum 2009).
Rotary heat exchangers generally provide the highest heat recovery efficiencies of any HRV or
ERV, and are less prone to fouling because air flowing in one direction removes particles
deposited by air flowing in the opposite direction. Rotary heat exchangers may be more
expensive to install and maintain because they require a lot of room, having moving parts, and
require a separate motor and control to provide rotation. Cross contamination of the airstreams is
also a concern, although a purge section can be used to reduce the amount of exhaust air entering
the supply air stream (Ruan et al. 2012).
2.3.3 Run-Around
Run-around heat exchangers are another form of regenerative heat recovery. Energy is
transferred from one air stream to a water-glycol mix, which is circulated by a pump to the other
airstream where the energy is released. Although less efficient than plate or rotary heat
exchangers, run-around heat exchangers have the advantage of being able to transfer energy
between two airstreams which may not be in close physical proximity. This can be particularly
important in building retrofits where the existing ventilation system was not designed with heat
recovery in mind. Another advantage of run-around systems over rotary heat exchangers is that
there is no airstream cross-leakage and, other than the pump, there are no moving parts.
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Although traditional run-around systems were limited to sensible energy recovery, some
researchers (Seyed-Ahmadi et al. 2009) have proposed using semi-permeable membranes in
combination with an aqueous salt solution to transfer both sensible and latent energy between the
two airstreams. A heat pump can be added to run-around systems to increase heat recovery
efficiency from 50% to nearly 70%, although installation costs and electricity consumption are
also increased (Wallin et al. 2012).
2.3.4 Heat Pipe
Heat pipes are growing in popularity for ventilation air dehumidification and reheating (Zhang
and Lee 2011). They are not commonly used for recovering energy from exhaust ventilation air
however, because they are limited to sensible energy only and are more expensive than other
types of heat exchangers. As outlined by Shao and Riffat (1997), heat pipes may be useful for
applications where a very low pressure drop is desired, such as with natural ventilation systems.
2.3.5 Alternating Flow Regenerators
Alternating flow regenerative HRVs function by moving air through an energy storage medium
in alternating directions. The energy storage medium stays in a fixed position, and the direction
of airflow alternates to capture energy from one airstream, and then deposit it into the other
airstream. Because each unit allows for airflow in a single direction, an even number of
synchronized ventilation units are required to achieve balanced airflow. There exist at least two
commercial variations of the reversible regenerator, one with an aluminum sheet energy storage
medium (Manz et al. 2000) and one with a ceramic energy storage medium (Schmidt and Klein
2011). To the authors’ knowledge there have been no peer-reviewed studies of the newer product
with ceramic storage medium.
For both products, the cycle lengths between changes in airflow direction are in the range of 75
to 80 seconds. Sensible efficiencies appear to be comparable to those achieved by plate heat
exchangers, and it is possible to recover latent energy, depending on the storage material.
Computational fluid dynamics models by Manz et al. (2000) indicate air change efficiencies of
0.63-0.83 (1.0 corresponds to complete mixing), depending on the location of the units relative to
each other. The air change efficiencies measured in this study suggest significant short circuiting
of fresh supply airstream back into the exhaust airstream, which may reduce overall ventilation
effectiveness.
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3 The Fine Wire HRV
The review of existing heat recovery technologies indicates that further study of decentralized
ventilation systems with heat recovery would be beneficial. The fine wire HRV represents a
novel concept for decentralized ventilation with heat recovery. The ventilator uses copper wires
to transfer sensible energy between the supply and exhaust airstreams. Although the heat
exchanger technology was first patented in North America in 1998 (Van Andel), it has yet to be
successfully commercialized. This section describes the heat recovery technology in more detail
and outlines a potential decentralized ventilation strategy for residential buildings.
Figure 4. Fine wire heat exchanger and manifold
3.1 Heat Recovery Technology
Manufactured by in the Netherlands Fiwihex (2006), the fine wire heat exchanger is composed
of 28 layers of 0.1 mm diameter copper wire. The wires are spaced at 0.6 mm parallel to the
direction of airflow, and 0.4 mm perpendicular to the direction of airflow. Adjacent air streams
are separated by 2 mm thick layers of glue. The copper wires extend between supply and exhaust
airstreams, through the glue, allowing for heat transfer via conduction. A thin lacquer protects
the copper from corrosion without significantly limiting heat transfer.
The copper wires form a matt roughly 454 mm long, 217 mm wide and 16 mm thick, which is
divided into 17 channels along the width (Figure 5a). The heat exchanger area is roughly 0.05 m2
in each direction of airflow. A single fine wire heat exchanger is composed of approximately
30,000 copper wires with a combined mass of roughly 500 g.
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The narrow diameter of the copper wire reduces the thickness of the boundary layer of air
formed at the surface, resulting in a high heat transfer coefficient between the copper and the air.
Because the copper wires are very conductive, efficient heat transfer between adjacent airstreams
is achieved, as illustrated in Figure 5b.
(a) Elevation (b) Section with details
Figure 5. Heat exchanger and heat recovery process
3.2 Enclosure and Manifold Design
An enclosure was built for the heat exchanger out of a combination of wood and transparent
acrylic plastic, with clear caulking and sheathing housewrap tape used for air sealing. Two
centrifugal fans were mounted back-to-back at the top of the enclosure, as shown in Figure 6.
The thickness of the enclosure (roughly 150 mm) was dictated by the depth of the fans. At this
thickness, the enclosure fits within the width of most conventional wall systems. For testing
purposes, the speed of each fan was controlled using a potentiometer and standard resistors to
create an adjustable 0-10 V output. Additional photos of the fans, control, enclosure and
manifold can be found in Appendix A.
High efficiency fans with brushless direct current (DC) motors were used. Electronically
commutated motors (ECMs) are brushless DC motors with an internal rectifier to convert
alternating current (AC) to DC. ECMs do not require an external DC power source, making them
popular for residential HVAC applications. For the purposes of testing, two 100 W variable
speed fans with non-electronically commutated brushless DC motors were used, requiring two
48 VDC power supplies.
Pre-Heated
Supply Air
Warm
Indoor Air
Cooled
Exhaust Air
Cold
Outdoor Air
28 L
ayers o
f Co
pp
er
Wire (ø
=0
.1m
m)
16mm
13mm Glue Boundary
Heat Transfer
via Conduction
454mm
217mm
Section Cut
14
Figure 6. Photos of variable speed brushless DC centrifugal fans and air sealing
A manifold was supplied with the heat exchanger provided by Fiwihex. A manifold is required
on both sides to keep the two airstreams separate as they enter and exit the heat exchanger. The
design is tapered such that the height is reduced at the top and bottom of the heat exchanger,
which likely reduces the airflow rate and alters the airflow direction the airflow at these
locations. With a maximum centre height of 40 mm, the manifold easily fits within the enclosure
depth. The enclosure and manifold are illustrated in Figure 7.
(a) Enclosure elevation (b) Enclosure section (c) Manifold airflow detail
Figure 7. HRV enclosure and heat exchanger manifold
Outlet
Copper Wires
Manifold
From Inlet
Centrifugal
Fans
Manifold
15
3.3 Distribution System
By integrating the fine wire HRV ventilation into the façade to create a “breathing wall”, the
extensive ductwork and space requirements of conventional ventilation systems can be avoided.
By using a system of individual ventilation units in each room, rather than a single centralized
unit, air can be distributed more efficiency and airflow to each room can be controlled
independently. This independent control is particularly beneficial when individual rooms are
unoccupied for extended periods, as it allows ventilation to be reduced without increasing
occupant pollutant exposure. Ventilation rates can also be increased if higher than normal
occupancy is detected in a room, better controlling occupant-generated pollutant levels and
improving indoor air quality. Although ventilation noise levels may be higher because of the
proximity of the HRV to occupied spaces, eliminating central air distribution ductwork can
reduce sound transmission between rooms. A typical installation of the fine wire HRV in a
bedroom is illustrated in Figure 8.
Figure 8. Typical use of the fine wire HRV in a bedroom
Balanced airflow to each room is recommended to maximize overall ventilation heat recovery.
Balanced airflow also helps prevent air movement through the building envelope, which can
result in moisture issues depending on the relative indoor and outdoor temperature and relative
humidity. A separate occupant-controlled exhaust fan is recommended for the bathroom and
16
kitchen range, as the high levels of moisture and pollutants at these locations would increase the
HRV cleaning and maintenance requirements. Make up air for the bathroom and kitchen is
provided through infiltration, or through temporary imbalances in airflow at the HRVs in
adjacent spaces. This concept is illustrated in Figure 9.
Figure 9. Fine wire HRV concept and distribution system
3.4 Ventilation Control
The current design for the heat exchanger assembly utilizes two variable speed fans to allow
independent control of the supply and exhaust airstreams. Although for testing purposes the fans
were controlled using a simple potentiometer, a site installation would benefit from a more
sophisticated programmable logic controller with pressure and temperature sensors in each
airstream. Using this control configuration, the speed of one fan can be modulated to maintain
equal pressure drop across the heat exchanger for both air streams, resulting in balanced supply
and exhaust airflow rates, even if the space is positively or negatively pressurized due to stack
effect, wind or the operation of other mechanical systems. By automatically balancing the two
airflows, overall heat recovery efficiency can be maximized.
Since the fine wire HRV can be integrated into the wall and used to provide decentralized
ventilation, it is ideally suited to DCV. In a residential setting, a preset schedule will provide
good control for rooms with relatively predictable occupancy patterns, such as bedrooms. The
addition of a manual override or PIR sensor may be helpful to activate the ventilation system if
the room is occupied outside of the normal schedule. In other areas of the home where the
number of people and their schedules are less predictable, such as the living room, PIR
occupancy sensors should be used in combination with CO2 sensors. This allows the ventilation
system to be activated as soon as the space is occupied, but also allows the system to estimate
17
occupancy levels, and increase ventilation accordingly. The ventilation control strategy is
explored further in Section 6.4.2. Further refinement of the occupancy detection and control
strategy requires study of an installed prototype in an occupied home.
Optionally, reversible fans can be used to allow for cross ventilation and free cooling during
favourable outdoor conditions. In this configuration, HRVs on the windward side of the building
would be set to supply air only, with one fan operating in reverse such that both fans are
pressuring the space. On the leeward side of the building the HRVs would be operating in a
similar manner, but would be depressurizing the space. Working in combination, significant
cross ventilation could be achieved. This increased ventilation would be beneficial when there is
a desire for cooling and outdoor temperatures are lower than indoor temperatures.
4 Performance Measurements
Third-party performance data for the fine wire HRV is not published by the manufacturer and to
the author’s knowledge there have been no peer-reviewed studies of the technology. To address
this gap, laboratory tests were conducted to assess performance in a number of areas. Efforts
were made to quantify the pressure drop, DC fan electricity consumption, airstream cross-flow
leakage and sensible heat recovery efficiency at varying flow rates.
4.1 Pressure Drop
A system curve relating static pressure drop to flow rate was developed for the heat exchanger.
The system curve is an important parameter when selecting fans and can be a significant factor in
fan electricity consumption. Knowing the system curve also simplifies future testing by allowing
airflow rates to be calculated indirectly from static pressure measurements. Static pressure drop
and flow rate were measured using a calibrated axial fan in combination with a digital
differential pressure and flow gauge. To simplify the process, the fine wire HRV’s centrifugal
fans were not operated during the test, and the calibrated axial fan was used to generate airflow.
Static pressure probes were inserted through the side of the enclosure and located away from the
surface of the heat exchanger in an area of low air velocity. The test setup, equipment and
location of pressure sensors are shown in Figure 10.
18
Figure 10. Test setup for pressure and flow rate measurements
The system curve can be expressed using Equation 1 (adapted from ASHRAE 2013).
(1)
where,
Q is the airflow rate through the heat exchanger [L/s]
C is the flow coefficient
ΔP is the static pressure difference across the heat exchanger [Pa]
n is the flow exponent
Measurements of static pressure drop were made at flow rates ranging from 5 L/s to 50 L/s. By
taking natural logarithms of both sets of values and doing a linear regression, the flow coefficient
and flow exponent could be estimated as shown in Figure 11a. Using these values, the system
curve could be generated, as shown in Figure 11b.
19
(a) Estimating flow coefficient and exponent (a) Measured values and system curve
Figure 11. Pressure and flow rate measurements used to develop system curve
The difference in static pressure drop measured across the two sides of the heat exchanger at a
given flow rate was less than 5% and the system curve shown in Figure 11b was considered
accurate for both sides.
4.2 Fan Electricity Consumption
HRV electricity consumption is an important factor in overall ventilation energy use. Some
North American HRVs make use of inefficient permanent split capacity (PSC) motors which
result in excessive fan electricity consumption, although the industry is moving towards more
efficient fans with brushless DC motors (Straube 2009). More efficient air-to-air heat exchangers
generally result in larger pressure losses and require more fan power to move the same volume of
air. As a result, there is often a tradeoff between higher heat recovery efficiency and lower fan
electricity consumption. Depending on the climate, the energy benefit of selecting a more
expensive HRV with higher heat recovery efficiency may be wholly or partially offset by
increased fan power (El Fouih et al. 2012).
The electrical power consumption of the fine wire HRV was measured using an outlet electricity
meter. Static pressure drop across the heat exchanger was measured and converted to airflow rate
using Equation 11. Power consumption measurements were taken with both fans operating, with
one fan operating and with both fans off as shown in Figure 12.
R² = 0.9988
0
1
2
3
4
5
6
0 1 2 3 4 5
ln(Δ
P)
ln(Q)
ln(ΔP) = 1.649 ln(Q) - 1.0638ln(ΔP) = ln(Q)/0.606 - 0.649/0.606ΔP = (Q/1.91)1/0.61
0
50
100
150
200
250
0 10 20 30 40 50
Stat
ic P
ress
ure
Dro
p, Δ
P(P
a)
Airflow Rate, Q (L/s)
Measured Values
System Curve
ΔP = (Q/1.91)1/0.61
20
Figure 12. Fine wire HRV electrical power consumption at different flow rates
Comparing electricity consumption of the fine wire HRV with one and two fans operating
suggests that there is an overhead power supply electricity consumption of about 12 W,
regardless of fan speed. At low flow rates, this is as much as 65% of total electricity
consumption. Considerable electricity savings may be possible by using a smaller power supply,
although this would limit the HRV’s maximum airflow capacity. If the fine wire HRV is to be
operated intermittently the control should cut power to the power supply when the fans are not
operating. A separate power source would be required for the control components.
Total electrical power per unit flow rate was calculated to illustrate ventilation efficiency at
varying flow rates and allow for comparison to conventional HRVs. The results are illustrated in
Figure 13.
0
20
40
60
80
100
120
140
160
10 20 30 40 50 60 70
Ele
ctri
cal
Po
we
r C
on
sum
pti
on
(W
)
Airflow Rate in Each Direction (L/s)
2 Fans Operating
1 Fan Operating
Power Supply Only
21
Figure 13. Fine wire HRV ventilation electrical efficiency at different flow rates
Power consumption per unit flow rate is minimized to 1.1 W per L/s at flow rates around 25 L/s.
This results from the overhead electricity consumption of the power supply, which has a greater
relative penalty at low flow rates. Using a smaller power supply would shift the optimal
operating point to a lower flow rate.
Fan electricity consumption of the fine wire HRV was compared to conventional HRVs listed in
the Home Ventilation Institute Certified Products Directory (2014). Where performance data was
given at multiple flow rates, the fan electricity consumption at the highest sensible heat recovery
efficiency was used. The mean electricity consumption of the conventional HRVs was 2.2 W per
L/s. At 1.1 W per L/s, only the top 5% of conventional HRVs consume less electricity than the
fine wire HRV. The distribution of conventional units is shown in Figure 14.
Figure 14. Distribution of electrical power consumption of HRVs
0.0
0.5
1.0
1.5
2.0
2.5
3.0
10 20 30 40 50 60 70
Ele
ctri
cal
Po
we
r C
on
sum
pti
on
(W p
er
L/s)
Airflow Rate in each Direction (L/s)
0%
10%
20%
30%
40%
50%
<=1.0 1.5 2.0 2.5 3.0 3.5 4.0 4.5 >4.5
Pe
rce
nta
ge o
f C
on
ven
tio
nal
HR
V U
nit
s
Electrical Power Consumption (W per L/s)
Fine WireHRV1.1 W per L/s
Source: Home Ventilation Institute 2014
22
It is important to note that the distribution shown in Figure 14 is based on all HRVs certified by
the Home Ventilation Institute and is not adjusted to account for the number of units sold. As
such, the distribution of HRVs in use in North American homes may differ.
4.3 Airstream Cross-Leakage
The Canadian Standards Association (CSA) standard C439, Standard laboratory methods of test
for rating the performance of heat/energy-recovery ventilators (2010) requires cross-leakage
between the exhaust and supply airstreams be measured using a tracer gas. Although a particular
tracer gas is not specified, the standard states that the tracer gas injection rate must be sufficient
such that a cross-leakage of 0.1% is within the range of the measurement device being used.
Sherman (1990) identified five ideal characteristics of tracer gases for determining ventilation
rates: safety, non-reactivity, insensibility, uniqueness and measurability. While early researchers
experimented with many different tracer gases, sulfur hexafluoride (SF6) has become the most
popular choice in recent years. While SF6 exhibits the five ideal characteristics defined by
Sherman and functions well as a tracer gas, it also has an extremely high global warming
potential, causing many to question the ethics of its use in research and leading some
jurisdictions to control its availability. As the use of SF6 becomes increasingly problematic,
researchers are exploring more environmentally friendly tracer gases (Burke et al. 2014).
For the purposes of this study, airstream cross-leakage testing was performed using CO2 as a
tracer gas. The decision was made based on the accessibility of CO2 (in the form of dry ice) and
CO2 sensors. While CO2 is not unique, background concentrations in the test room were
considered stable enough that they would not have a significant impact on the test results. The
measurability of CO2 using the sensors available also presented a challenge, both in terms of
sensor noise and range, but these issues were managed as described in Section 4.3.3.
4.3.1 Test Set Up and Methodology
Three CO2 sensors were installed as shown in Figure 15. CO2 was produced by crushing dry ice
and placing it within a capture hood, through which Airstream A was circulated. Sublimation of
the dry ice quickly increased the CO2 concentration within Airstream A, which was recirculated
through the heat exchanger in a closed loop. The CO2 concentration in Airstream A was
measured on the inlet side of the heat exchanger (CO2 Sensor 1). At the same time, ambient air
23
was circulated through Airstream B, with CO2 concentrations being measured on the inlet side
(CO2 Sensor 2) and outlet side (CO2 Sensor 3) of the heat exchanger. The elevated concentration
of CO2 in Airstream A was allowed to decay slowly while Airstream B was monitored for an
increase in CO2 concentration. An increase in CO2 concentration in Airstream B is indicative of
air leakage between the two airstreams.
Figure 15. Dry ice and configuration of leakage testing using CO2
Roughly 1-2 g of crushed dry ice was used to increase the CO2 concentration in Airstream A.
Since CO2 concentrations were measured directly, an accurate measurement of the mass of dry
ice used was unnecessary. The CO2 sensors used had a range of 0-5000 ppm and an accuracy of
+/- 50 ppm or +/- 3% of reading. The sensors were calibrated prior to the test using their internal
calibration function with outdoor air as a reference, and were recalibrated relative to each other
based on the ambient CO2 concentration measured during the test. A high degree of absolute
accuracy was unnecessary, as the test methodology relied on the relative CO2 concentrations
between the two airstreams. Measurements were taken every 10 seconds and converted to
average concentrations per minute to reduce the inherent CO2 sensor noise.
4.3.2 Measurements
The cross-leakage test was conducted at three airflow rates: 6.9 L/s, 11.7 L/s and 20.6 m3/h. The
measured CO2 concentrations are shown in Figure 16. Airflows were balanced using differential
pressure measurements across the heat exchanger in both airflow directions. Airflow rate was
calculated from differential pressure using the system curve developed in Section 4.1.
24
a) 6.9 L/s b) 11.7 L/s c) 20.6 L/s
Figure 16. CO2 concentrations in Airstreams A and B
The mass of dry ice used and the relatively small volume of the closed duct system mean that the
CO2 concentration in Airstream A quickly surpassed the upper limit of the sensor’s range (5000
ppm or approximately 4500 ppm above the ambient CO2 concentration). In all three tests, the
CO2 concentration in Airstream B showed a small but discernible increase correlating with the
sharp increase in CO2 concentration in Airstream A, indicating cross-leakage between the two
airstreams. While this increase was most pronounced at Sensor 3, there was also an observable
increase in the CO2 concentration measured at Sensor 2, as shown in Figure 17.
a) 6.9 L/s b) 11.7 L/s c) 20.6 L/s
Figure 17. CO2 concentrations in Airstream B before and after heat exchanger
0
1000
2000
3000
4000
5000
0 20 40 60
CO
2C
on
c. (
pp
m a
bo
ve a
mb
ien
t)
Time (minutes)
0
1000
2000
3000
4000
5000
0 20 40 60
CO
2C
on
c. (
pp
m a
bo
ve a
mb
ien
t)
Time (minutes)
0
1000
2000
3000
4000
5000
0 20
CO
2C
on
c. (
pp
m a
bo
ve a
mb
ien
t)
Time (minutes)
0
50
100
150
200
250
0 20 40 60
CO
2C
on
cen
trat
ion
(pp
m a
bo
ve a
mb
ien
t)
Time (minutes)
0
50
100
150
200
250
0 20 40 60
CO
2C
on
cen
trat
ion
(pp
m a
bo
ve a
mb
ien
t)
Time (minutes)
0
50
100
150
200
250
0 20
CO
2C
on
cen
trat
ion
(pp
m a
bo
ve a
mb
ien
t)
Time (minutes)
25
Because Sensor 2 was located on the inlet side of the heat exchanger, the increase in CO2
concentration measured at this location is not a result of air leakage through the heat exchanger
itself. The most probable explanation is that air leakage occurred elsewhere in the enclosure
before Sensor 2, likely at the bases of the centrifugal fans which penetrate the surface dividing
Airstreams A and B. Although it is also possible that the ambient CO2 concentration in the test
room became elevated as a result of CO2 produced during the experiment, it is unlikely given the
large volume of the room and relatively small mass of CO2 used.
4.3.3 Estimating Cross-Leakage
Although leakage through the heat exchanger is most accurately represented by the increase in
CO2 concentration between Sensor 2 and Sensor 3, comparing CO2 concentration at Sensor 3 to
ambient levels provides an estimate of cross-leakage through the complete enclosure and heat
exchanger. The higher CO2 concentration at Sensor 3 also makes the trend easier to distinguish
against the background noise. For this reason, CO2 concentration in Airstream B was based on
the increase in CO2 concentration above ambient levels at Sensor 3.
Because the CO2 concentration in Airstream A is outside the sensor’s range during the period in
which there is an observable increase in CO2 concentration in Airstream B, the CO2
concentration in Airstream A must be estimated. This can be accomplished using Equation 2a
(adapted from ASHRAE 2013).
(2a)
where,
t time from the start of the test [min]
C(t) CO2 concentration above the ambient concentration at time t [ppm]
C0 CO2 concentration above the ambient concentration at time zero [ppm]
S CO2 source rate [ppm/min]
L CO2 loss rate [min-1
]
By assuming that the dry ice sublimates quickly enough to result in a net increase in CO2
concentration in Airstream A throughout the sublimation process, it follows that the peak CO2
concentration (Cpeak) in Airstream A coincides with the time at which sublimation is complete.
After the peak time (tpeak), the CO2 source rate, S, is zero, reducing Equation 2a to Equation 2b.
26
(2b)
Although the point of peak CO2 concentration for Airstream A is unknown because the
concentration is above the sensor’s upper limit, it corresponds with the point of peak CO2
concentration in Airstream B, which is within the sensor’s limits. The CO2 loss rate, L, can be
determined from the declining period of CO2 concentration (after the peak time), which is only
measurable once concentrations are below the sensor’s upper limit. The CO2 concentration in
Airstream A can then be estimated for the period after sublimation of the dry ice is complete and
before concentrations are within the range of Sensor 1. This process is illustrated in Figure 18 for
the test conducted at 11.7 L/s.
Figure 18. Process for estimating CO2 concentrations in Airstream A
Following the same process for the other two tested airflow rates, CO2 concentration in
Airstream A can be approximated and compared to CO2 concentrations in Airstream B to
estimate air leakage using Equation 3 (adapted from Canadian Standards Association 2010).
(3)
Using the estimated CO2 concentration in Airstream A and the measured CO2 concentration
Airstream B, the cross-leakage rate for the three tested airflows varied from 0.5% to 1.5%.
0
200
400
600
800
1000
1200
1400
1600
1800
2000
0
2000
4000
6000
8000
10000
12000
14000
16000
18000
20000
0 10 20 30 40 50
CO
2C
on
c. -
Air
stre
am B
(p
pm
ab
ove
am
bie
nt)
CO
2C
on
c. -
Air
stre
am A
(p
pm
ab
ove
am
bie
nt)
Time (minutes)
Airstream A (Sensor 1)
Airstream A (estimated)
Airstream B (Sensor 3)
tpeak = 11 miny = -0.1403x + 9.8123
R² = 0.9954
L = 0.1403 min-1
Cpeak = e9.8123 = 18256 ppm
2.0
3.0
4.0
5.0
6.0
7.0
8.0
9.0
10 20 30 40 50
ln (C
O2
con
cen
trat
ion
)
t - tpeak (minutes)
27
4.3.4 Limitations
The results obtained for air leakage between the two airstream are estimates, limited by the noise
and range of the CO2 sensors. In addition, Equations 2a and 2b assume uniform CO2
concentration within each airstream, which may not be true, particularly if air leakage through
the heat exchanger is localized. More accurate measurements could be obtained by using a
combination of tracer gas and multiple sensors with increased sensitivity and range. While
estimates, these preliminary results do indicate that the fine wire heat exchanger and enclosure
are relatively air tight, and that cross contamination of the two airstreams is unlikely to have a
significant impact on energy efficiency or indoor air quality. It can also be concluded that air
leakage between the two airstreams is not a major factor affecting heat recovery efficiency.
4.4 Sensible Heat Recovery Efficiency
In cold climates, the main benefit of installing a more expensive balanced HRV system over a
lower cost exhaust-only ventilation system is the capacity to recover heat and reduce the
ventilation heating load. It follows that the heat recovery efficiency of an HRV is an important
parameter in assessing its overall value and is often the first performance characteristic
considered when selecting an HRV.
4.4.1 Measurement Standards
There are a number of ways of measuring sensible heat recovery efficiency. The compact design
of most HRVs requires that temperatures be measured at the main inlet and outlet ducts, with the
fans located between the points of measurement. As a result, heat generated by the fans increases
measured temperatures, often inflating the apparent efficiency. Most reputable measurement
standards correct for heat generated by fans and other factors which may distort the results.
The North America-based Heating Ventilation Institute publishes results based on CSA C439,
Standard laboratory methods of test for rating the performance of heat/energy-recovery
ventilators (2010). Sensible heat recovery efficiency is calculated using Equation 4.
(4)
28
where,
n total number of measurements
i ith
time that data are recorded
Ms net mass flow rate of the supply air, after accounting for cross-leakage [kg/s]
Cp specific heat of the air [kJ/kg°C]
t5 net temperature at the supply air outlet, after correcting for cross-leakage [°C]
t1 dry-bulb temperature of supply air inlet [°C]
t3 dry-bulb temperature of exhaust air inlet [°C]
Δϴ time between flow measurements [s]
QSF energy input into supply airstream attributed to fan(s) [kJ]
QSH energy used by heater and compressor in supply airstream [kJ]
QC casing heat transfer [kJ]
QD energy used for defrost [kJ]
QL heat loss due to casing leakage [kJ]
Mmax greater of the net mass flow rate of the exhaust and supply air [kg/s]
QEF energy input into exhaust airstream attributed to fan(s) [kJ]
QEH energy used by heater in exhaust airstream [kJ]
CSA C439 also defines an additional parameter called apparent sensible effectiveness, which
does not correct for fan heat or cross-leakage. As a result, it is typically higher than the sensible
heat recovery efficiency, particularly for HRVs with inefficient fans. Although of less technical
relevance than sensible heat recovery efficiency, apparent sensible effectiveness may be quoted
by manufacturers to increase the perceived performance of their product.
The German-based Passive House Institute has also developed a standard for calculating sensible
heat recovery efficiency, using Equation 5 (Passivhaus Institut 2009).
(5)
where,
Pel exhaust air fan electrical power [W]
ṁ mass flow rate of the exhaust air [kg/s]
Cp specific heat of the air [kJ/kg°C]
t1 dry-bulb temperature of supply air inlet [°C]
t3 dry-bulb temperature of exhaust air inlet [°C]
t4 dry-bulb temperature of exhaust air outlet [°C]
Equation 5 assumes balanced flow rates and negligeable air leakage. The correction factor for
fan energy also makes the assumption that the supply air fan is located on the outlet side of the
heat exchanger, as shown in Figure 19. Equation 5 will slightly under calculate sensible heat
recovery efficiency if the supply air fan is located on the inlet side of the heat exchanger.
29
Figure 19. Location of temperature measurements in standard test procedures
4.4.2 Test Set Up and Methodology
The sensible heat recovery efficiency of the fine wire HRV was measured in an indoor
laboratory using the setup illustrated in Figure 20. A calibrated axial fan was used to generate the
supply airflow, while a variable speed centrifugal fan was used to generate the exhaust airflow.
Static pressures were measured on the inlet and outlet sides of the heat exchanger in both
airstreams, at the same locations used to generate the system curve in Section 4.1. Temperature
measurements were taken at 10 second intervals with sets of three Type T thermocouples
(resolution 0.1 °C) located at each inlet and outlet.
Indoor air at approximately 21 °C was used for the supply airstream. Two 200 W incandescent
light bulbs connected to a dimmer switch were used to heat the exhaust airstream, creating a
temperature difference of 10-20 °C between the two inlets. The measurements described in
Section 4.1 indicate that pressure drop across the two sides of heat exchanger is approximately
equal for a given airflow rate. Based on this observation, airflow rates were balanced by
equalizing the pressure drop across the heat exchanger and airflow rate was recorded using the
calibrated axial fan located at the supply air inlet.
Indoors
Exhaust Inlet
t1
t2 (t5)
t3
t4
Outdoors
Exhaust Outlet
Supply Inlet
Supply Outlet
30
(a) Test setup and airflows (b) Location of thermocouples
Figure 20. Heat recovery efficiency test setup and location of thermocouples
Since it was possible to take temperature measurements between the heat exchanger and the fans,
it was not necessary to correct for waste heat. Cross-leakage was assumed negligible based on
the tracer gas measurements described in Section 4.3. Heat transfer through the HRV enclosure
was also assumed to have a negligible effect given the close proximity of the temperature
measurements to the heat exchanger, as well as the low temperature differences between the
airstreams and the ambient air in the laboratory. Under these conditions Equation 4 from CSA
C439 and Equation 5 from the Passive House Institute can be reduced to just the temperature
variables. The numerator is different in the two equations, with Equation 4 using the increase in
temperature of the supply airstream and Equation 5 using the decrease in temperature of the
exhaust airstream. If airflows are balanced, the reduced forms of the two equations give the same
result. To achieve the most reliable results and account for any airflow imbalances during testing,
the increase in temperature of the supply airstream and decrease in temperature of the exhaust
airstream were averaged using Equation 6 (proposed by the author based on Equations 4 & 5).
(6)
where,
t1 average dry-bulb temperature of supply air inlet [°C]
t2 average dry-bulb temperature of supply air outlet [°C]
t3 average dry-bulb temperature of exhaust air inlet [°C]
t4 average dry-bulb temperature of exhaust air outlet [°C]
Cardboard Warming Hood
Calibrated Axial FanCentrifugal Fan
Exhaust Inlet Supply Inlet
Supply Outlet Exhaust Outlet
Heat exchanger and manifold
Supply InletExhaust Inlet
Supply Outlet
Exhaust Outlet
t1t3
Thermocouples
t4t2
31
4.4.3 Results
A series of nine tests was conducted to measure the sensible heat recovery efficiency of the fine
wire HRV at airflow rates ranging from 5.7 L/s to 26.4 L/s. The heat output at the exhaust inlet
was increased with increasing flow rate to maintain a temperature rise of 10-20 °C over the
supply inlet. Temperatures were averaged over a 5-10 min period after they had stabilized. The
results for airflow rates of 9.7 L/s and 18.4 L/s are shown in Figure 21. The complete set of
temperature measurements can be found in Appendix B.
(a) Airflow rate of 9.7 L/s (b) Airflow rate of 18.4 L/s
Figure 21. Temperature measurements at two airflow rates
The temperature measurements consistently show a larger decrease in temperature of the exhaust
airstream than increase in temperature of the supply airstream. The difference is greatest at low
airflow rates and gradually decreases with increasing airflow rate. It is unknown exactly why this
difference occurred. The measurements described in Section 4.3 indicate that cross-leakage is
unlikely to be a significant factor. Temperature measurements were taken with sets of three
thermocouples at each location. The difference between thermocouple readings does not suggest
significant variation across the airstream section or malfunctioning sensors.
Despite the measurements described in Section 4.1 which indicate very similar system curves for
both sides of the heat exchanger, it is possible that the geometry of the exhaust side creates a
higher pressure drop, resulting in a slightly lower exhaust airflow rate. Air density may also be a
factor, as the exhaust airstream is warmer and less dense, resulting in a slightly lower mass flow
20
22
24
26
28
30
32
34
36
38
40
0 100 200 300 400
Me
asu
red
Te
mp
era
ture
(°C)
Time (seconds)
t3
t2
t4
t1
ExhaustΔt = 15.1 °C
SupplyΔt = 12.1 °C
20
22
24
26
28
30
32
34
36
38
40
0 50 100 150 200 250 300
Me
asu
red
Te
mp
era
ture
(°C)
Time (seconds)
t3
t2
t4
t1
ExhaustΔt = 11.2 °C
SupplyΔt = 9.8 °C
32
rate at equal volume flow rates. Based on conservation of energy, unequal mass flow rates would
result in different temperature changes for the two airstreams.
Another possibility is that heat transfer through the enclosure is causing a reduction in airstream
temperature as it travels between the heat exchanger and the thermocouples. Since the exhaust
outlet and supply inlet airstreams are roughly the same temperature as the ambient air, any heat
transfer through the enclosure at these locations should be negligible. In contrast, both the
exhaust inlet and supply outlet airstreams are 10-20 °C warmer than the ambient air and
significant heat loss through the enclosure is possible, as shown in Figure 22.
Figure 22. Effect of heat loss through the HRV enclosure
Assuming a surface film conductance of 8 W/m2/K for the exterior of the enclosure, and
negligible thermal resistance for the interior surface film (high air velocity) and enclosure itself,
as much as 22 W of heat could be lost from each of the exhaust inlet and supply outlet
airstreams. At an airflow rate of 9.7 L/s this heat loss could result in a 2 °C change in
temperature in each airstream, which is sufficient to explain the discrepancy shown in Figure
21a. It would also explain why the magnitude of the discrepancy decreases with increasing
airflow rate, as the same rate of heat loss is being distributed over a larger volume of air.
Supply InletExhaust Inlet
Supply Outlet
Exhaust Outlet
t1t3
Thermocouples
t4t2
Small Δt between airstreams and ambient air results in negligible
heat loss
Large Δt between airstreams and ambient air results in significant
heat loss
33
A preliminary attempt to test this theory was made by thermally insulating the enclosure and
repeating the test, but a difference in temperature changes of the two airstreams was still
observed. As the test was done in haste and without careful documentation, the results are
considered inconclusive. Future testing should be done using an insulated enclosure to minimize
heat transfer.
The potential for bias introduced by unbalanced mass flow rates is addressed by using Equation
6 to average the decrease in temperature of the exhaust airstream with the increase in
temperature of the supply airstream. This approach also addresses the issue of heat loss through
the enclosure, assuming that the heat lost in the exhaust inlet and supply outlet airstreams is
approximately equal. More reliable results could be obtained by using two calibrated fans to
more accurately balance airflow rates and by better insulating the HRV enclosure during testing
to reduce heat loss. The resulting sensible heat recovery efficiencies are shown in Figure 23.
Figure 23. Sensible heat recovery efficiency test results
A maximum sensible heat recovery efficiency of 82% was achieved at airflow rates of 7-10 L/s.
Efficiency declined linearly at higher airflow rates, with each 2 L/s increase in airflow resulting
in a 1% reduction in sensible heat recovery efficiency.
0.0 0.1 0.2 0.3 0.4 0.5 0.6
70%
75%
80%
85%
90%
0.0 5.0 10.0 15.0 20.0 25.0 30.0
Average Surface Air Velocity (m/s)
Sen
sib
le H
eat
Re
cove
ry E
ffic
ien
cy
Airflow Rate (L/s)
34
The maximum sensible heat recovery efficiency of the fine wire HRV was compared to
conventional HRVs listed in the Home Ventilation Institute Certified Products Directory (2014).
Where the directory provides performance data at multiple flow rates, the highest value was
used. The mean sensible heat recovery efficiency of the conventional HRVs is 68%. Only 2.5%
of conventional HRVs have higher sensible efficiencies than the fine wire HRV. The distribution
is shown in Figure 24.
Figure 24. Distribution of sensible heat recovery efficiencies of HRVs
The distribution shown in Figure 24 is based on all HRVs certified by the Home Ventilation
Institute and is not adjusted to account for the number of units sold. As such, the distribution of
HRVs in use in North American homes may differ.
5 Preliminary CFD Model
As part of this study of the fine wire heat exchanger, a preliminary investigation using
computational fluid dynamics (CFD) modeling was started. CFD uses equations of fluid
dynamics in combination with numerical methods to predict fluid flows. OpenFOAM, an open-
source CFD modeling software, was used to model a single copper wire extending between two
adjacent airstreams, as shown in Figure 25. The model space was divided into a mesh, with
increasing numbers of mesh elements at the intersection of the different materials.
0%
10%
20%
30%
40%
55% 60% 65% 70% 75% 80% 85% 90% 95%
Pe
rce
nt
of
Co
nve
nti
on
al H
RV
Un
its
Maximum Sensible Heat Recovery Efficiency
Fine Wire HRV82%
Source: Home Ventilation Institute 2014
35
(a) Geometry used in model (b) Section showing temperatures and velocities
Figure 25. Initial CFD modeling results
A solver was selected which could calculate both turbulent flow of incompressible fluids, as well
as heat transfer between solid and fluid regions. In this way it was possible to model heat transfer
from one airstream to another via the copper wire. Had more time been available, the next step
would be to expand the model to include a full cross-section of 28 copper wires. This would
allow for parametric modeling to analyze the effect of different numbers, sizes and spacings of
copper wires on pressure drop and heat recovery efficiency of the heat exchanger.
CFD would also be a valuable tool for better understanding airflow through the heat exchanger
manifold. From the shape of the manifold, it seems that air velocities through the heat exchanger
are likely to be higher at the centre, and lower near the top and bottom. Since testing has revealed
different sensible heat recovery efficiencies at different total airflow rates, it seems likely that the
distribution of air velocities through the heat exchanger would influence its performance. CFD
could be used to analyze this effect in more detail.
Due to the many variables which can greatly influence the results of a CFD model, reliability can
only be achieved when physical measurements are used to calibrate and validate the model. One
challenge encountered in the early stages of CFD modeling of the fine wire heat exchanger was
the scale. With airstream channels of little more than 10 mm in width, it was difficult to
accurately measure air velocities using a standard hot-wire anemometer. It was also difficult to
locate the thermocouples in a specific location, as the size of the thermocouples was substantial
in comparison to the width of the channels. These challenges will need to be addressed in any
future modeling efforts.
Copper wire
Separating layer
36
6 Whole-Building Model
To evaluate the potential for whole-building energy savings, a model of ventilation energy usage
and indoor air quality was developed and used to test a variable ventilation strategy. The model
was based on a new energy-efficient three-bedroom home in Toronto. The variable ventilation
strategy consisted of four fine wire HRVs operating at adjustable ventilation rates in each zone
of the home, utilizing demand controlled ventilation principles. A conventional HRV operating
at a constant ventilation rate in the same home was also modeled for comparison purposes.
6.1 Indoor Air Quality Metrics
In proposing an alternative ventilation strategy, it is important to understand the impact on
indoor air quality. This was done by using a transient mass balance model to track concentrations
of three important residential pollutants: fine particulate matter less than 2.5 μm in aerodynamic
diameter (PM2.5), formaldehyde (HCHO), and ozone (O3). Based on research by Logue et al.
(2012), these three pollutants are among the six most costly, in terms of daily adjusted life years
(DALYs) lost, due to chronic inhalation in US residences. The three pollutants were chosen
based on their widespread presence in almost all homes and because sufficient measurement data
exists to model their behavior in a residential environment. They also present a balanced
combination of sources including predominately indoor (HCHO), predominately outdoor (O3),
and both indoor and outdoor (PM2.5). Other pollutants which top the list produced by Logue et
al., such as secondhand smoke and radon, also have considerable societal health costs but are
best approached by controlling the pollutant source, rather than through ventilation.
6.1.1 Fine Particulate Matter (PM2.5)
A number of studies have linked PM2.5 with adverse health effects. Dominici et al. (2006)
studied hospital emission rates and ambient PM2.5 levels in 204 urban communities throughout
the US and found statistically significant effects for many respiratory and cardiovascular
diseases. The greatest association was found for heart failure, which increased in risk by 1.28%
for every 10 μg/m3 increase in ambient PM2.5 levels. Franklin et al. (2007) studied the link
between ambient PM2.5 and reported deaths in 27 US communities, and found a 1.78% increase
in respiratory related mortality and a 1.03% increase in stroke related mortality for every 10
μg/m3 increase. There is no recognized threshold below which PM2.5 levels do not have health
37
effects and Health Canada (2014) recommends reducing concentrations as much as possible.
Outdoor PM2.5, which is generated by combustion (largely from electricity generation and
transportation), can be a significant source of indoor PM2.5. PM2.5 can also be generated indoors
directly by combustion activities such as smoking or cooking, by re-suspension of settled
particles and by reactions between other components of indoor air (Health Canada 2014).
6.1.2 Formaldehyde (HCHO)
HCHO is a volatile organic compound which has respiratory and allergic health effects and may
contribute to the development of asthma. Research of 148 children in 80 Australian houses by
Garrett et al. (1999) found a relationship between HCHO in homes and incidences of atopy, as
well as the severity of allergy symptoms. Other studies have found increases in nose, throat and
eye irritation with increasing indoor HCHO levels (Health Canada 2005). While there is
evidence that HCHO may cause cancer in the nasal cavity, these effects have only been observed
at concentrations higher than those found in most residential environments (Health Canada
2014). Health Canada (2014) sets a residential short-term (1 hr) HCHO exposure limit of 123
µg/m3 and a long-term (8 hr) exposure limit of 50 µg/m
3. In non-smoking households, the main
source of HCHO is off-gassing of formaldehyde-containing building materials and furnishing
such as particle board, paint and carpeting (Health Canada 2014). HCHO may also be produced
by reactions of other volatile organic compounds with O3 (Weschler et al. 1992).
6.1.3 Ozone (O3)
Ambient O3 has been linked to respiratory and cardiovascular health effects, including
exacerbation of asthma (Stephens et al. 2012). Controlled human exposure studies have shown
prolonged exposure to O3 at concentrations as low as 240 µg/m3 results in decreased lung
function and subjective respiratory symptoms (Health Canada 2014). O3 can also react with
substances in the indoor environment to generate volatile organic compounds, including HCHO,
ultrafine particles and other pollutants which further contribute to adverse health effects
(Weschler et al. 1992, Zhao et al. 2007). Health Canada (2014) sets a residential long-term (8 hr)
O3 exposure limit of 40 µg/m3. The main source of indoor O3 is from photochemical reactions
outdoors. People spend up to 90% of their time indoors, and a large portion of overall exposure
occurs within residences (Lee et al. 1999). Indoor sources of ozone were not considered in the
model as they are relatively rare in a residential environment.
38
6.2 Methodology: Energy Use
The study home was modeled with a floor area of 186 m2 based on the average single-family
detached house reported in a survey by the Canadian Home Builders’ Association (2013). The
house was assumed to have a ceiling height of 2.5 m and a volume of 465 m3. A whole-house
infiltration rate was chosen based on CONTAM software modeling of annual infiltration rates
for US houses conducted by Persily et al. (2010). The results of the study indicate a typical
infiltration air change rate of 0.09 hr-1
for the tightest 10% of single-family detached houses built
in 1990 or later. This value was used in the model and assumed to be a reasonable approximation
of infiltration rates for energy efficient housing constructed in Canada today.
The energy consumption of a conventional HRV operating at a constant ventilation rate was
compared to four fine wire HRVs operating at variable ventilation rates. The conventional HRV
model was based on a single ventilation zone. The sensible heat recovery efficiency and fan
electricity consumption were obtained from the Home Ventilation Institute Certified Products
Directory (2014). Values were averaged over models with net supply airflow rates between 30
L/s and 60 L/s (80 models). Where the directory provided multiple values at different airflow
rates for the same model, the values at airflow rates closest to 30 L/s were used. The mean
sensible heat recovery efficiency and fan electricity consumption of the sample were 61% and
1.99 W per L/s, respectively.
The fine wire HRV model was split into four ventilation zones: the main living space zone,
including an open concept kitchen and living room, and three separate bedroom zones. The main
living zone occupied half the total floor area and was ventilated by two fine wire HRVs. The
bedrooms each occupied a sixth of the floor area and were ventilated with a single fine wire
HRV. To maximize the benefits of decentralized ventilation and to simplify the model, it was
assumed that the partitions between zones were airtight and that bedroom doors were kept
closed, making air movement between the zones negligible. The infiltration rate was assumed to
be the same in each zone.
The heat recovery efficiency and fan electricity consumption of the fine wire HRV were based
on the measurements described in Section 4. In order to model variable airflow rates, linear
regressions with the measured data points were used to create equations relating airflow rate to
sensible heat recovery efficiency and fan electricity consumption. At the minimum airflow rate
39
of 15 L/s, a sensible heat recovery efficiency of 80% and a fan electricity consumption of 1.14 W
per L/s were used for the fine wire heat exchanger. For those periods when a lower ventilation
rate was desired, the unit was assumed to operate intermittently at the minimum airflow rate.
Fan electricity consumption was not adjusted to account for filters as data for conventional
HRVs is not published in the Home Ventilation Institute Certified Products Directory. Studies
have also shown that the energy cost of more efficient filters in residential and light-commercial
HVAC systems to be negligible (Stephens et al. 2010). The contribution of fan electricity to the
heating load was not considered. Although separate bathroom and kitchen exhaust ventilation is
recommended, they were assumed to operate intermittently and to contribute little to the overall
ventilation rate.
The impact of sensible heat recovery efficiency on annual ventilation heating load was assessed
using Equation 7 (adapted from ASHRAE 2013).
(7)
where,
Heat Loss is the annual heat loss due to mechanical ventilation [kJ]
q is the annual average ventilation rate [m3/day]
cP is the specific heat capacity of air [1.005 kJ/kg.°C]
ρ is the density of air [1.3 kg/m3]
HDD is the annual number of heating degree days [3734 °C.days]
η is the annual average sensible heat recovery efficiency
A value of 3734 heating degree days (HDD) per year was used, based on the October to May
degree days below 18 °C for Toronto Pearson International Airport measured from 1981 to 2010
(Environment Canada 2014). The impact of ventilation on energy for cooling was not considered
as the fine wire HRV is only capable of recovering sensible heat and is not well suited to
cooling-dominated climates.
6.3 Methodology: Indoor Air Quality
Indoor air quality was determined using a transient mass balance air pollutant model which
tracked concentrations of PM2.5, HCHO and O3. Pollutant concentrations were modeled at 5-
minute intervals over a 24-hour period using Equation 8 (adapted from ASHRAE 2013).
40
(8)
where,
C(t) is the pollutant concentration at the end of the time interval [μg/m3]
Co is the pollutant concentration at the beginning of the time interval [μg/m3]
t is the length of the time interval [hr]
L is the average pollutant loss rate over the time interval [hr-1
]
S is the average pollutant source rate over the time interval [μg/m3/hr]
The pollutant loss rate and source rate were calculated using Equation 9a and Equation 9b,
respectively (adapted from ASHRAE 2013).
(9a)
(9b)
where,
λinf is the infiltration rate [hr-1
]
λmech is the mechanical ventilation rate [hr-1
]
β is the pollutant deposition loss rate [hr-1
]
E is the indoor pollutant emission rate [μg/m3/hr]
Cout is the outdoor pollutant concentration [μg/m3]
P is the pollutant penetration factor
η is the filter efficiency
Parameter estimates were obtained from the literature and are discussed further in the following
sections. Except for the mechanical ventilation rate, the same parameter values were used in both
the conventional and variable ventilation models. The values are summarized in Table 1.
Table 1. Mass balance model parameter values
Parameter PM2.5 HCHO O3
Cout, Outdoor Concentration [μg/m3] 8.3 0 26.2
E, Indoor Emission Rate [μg/m3/hr] 10.3 18.2 0
β, Deposition Loss Rate [hr-1
]
With ventilation on 0.42 0 5.6
With ventilation off 0.21 0 2.8
P, Penetration Factor 0.72 NA 0.79
η, Filter Efficiency 19% NA 10%
The model assumes that each zone is completely mixed, and that there is no mixing between
zones. The model also assumes that all indoor sources are accounted for in the indoor emission
rates and that the deposition loss rates account for chemical reactions between pollutants.
41
6.3.1 Fine Particulate Matter (PM2.5)
Outdoor concentrations of PM2.5 were based on ambient measurements taken in downtown
Toronto by the Ontario Ministry of Environment and Climate Change (2014). A typical outdoor
level of 8.3 μg/m3 was used in the model based on the 2013 annual mean concentration.
Indoor source emission, deposition loss and penetration rates used for PM2.5 were based on the
research conducted by Williams et al. (2003). The study measured PM2.5 in 37 North Carolina
residences over a one-year period. Based on the indoor-generated PM2.5 concentrations,
deposition loss rates and air exchange rates measured in the study, an indoor source emission rate
of 10.3 μg/m3/hr was used in the model. The study found a penetration rate of 0.72, which agreed
well with the summer rate of 1.11 and winter rate of 0.54 measured by Long et al. (2001). A
PM2.5 penetration rate of 0.72 was used in the model.
The study by Williams et al. also found a mean deposition loss rate of 0.42 hr-1
. Of the
residences in the sample, 70.2% had central forced air mechanical systems and only reported
using natural ventilation 23% of the time. As a result, the PM2.5 deposition loss rate measured by
Williams et al. was used for periods when the HVAC system was running. Emmerich and
Nabinger (2001) found that deposition loss rates for PM2.5 were approximately halved when
ventilation systems were not running, and so a PM2.5 deposition loss rate of 0.21 hr-1
was used
for those periods. While other studies (eg. Long et al.) have found lower deposition loss rates for
PM2.5, they generally included fewer residences and shorter study periods. A standard filter
efficiency of 19% was used in the model, based on the value used by Riley et al. (2002) for
loaded residential furnace filters.
6.3.2 Formaldehyde (HCHO)
An indoor source emission rate of 18.2 μg/m3/hr was assumed for HCHO based on the mean
value of 14 US houses analyzed by Sherman and Hodgson (2004). This emission rate was
calculated from a steady-state mass balance as the product of the air exchange rate and indoor
minus outdoor concentration. As such, the emission rate reflects the net generation of HCHO
after accounting for emissions and deposition losses and an additional deposition loss rate was
not included in the model.
42
Following Sherman and Hodgson, a number of simplifying assumptions were made in modeling
HCHO concentrations. The emission rate was assumed to be independent of concentration,
which may over-predict HCHO levels at high concentrations as emissions tend to be lower under
these conditions. This assumption is necessary as the relationship between HCHO emission rate
and concentration varies by source and is not easily modeled. This assumption is also
conservative in the sense that it over-predicts the buildup of HCHO when the ventilation rate is
reduced, underestimating the benefits of a variable ventilation rate. As outdoor concentrations of
HCHO are generally much lower than indoor concentrations, they were assumed to be
negligible. This assumption eliminates the need for an HCHO penetration factor or filter removal
efficiency.
6.3.3 Ozone (O3)
Outdoor concentrations of O3 were based on ambient measurements taken in downtown Toronto
by the Ontario Ministry of Environment and Climate Change. A typical outdoor level of 26.2
μg/m3 was used in the model based on the 2013 annual mean concentration. While some indoor
sources of O3 do exist such as air cleaners and photocopiers, they are uncommon in residential
settings and were not considered in the model. Therefore, a negligible O3 indoor source emission
rate was assumed.
Lee et al. measured O3 deposition loss rates in 43 Southern California homes and recorded a
mean value of 2.8 hr-1
. Nearly all of the homes included in the study were single family detached
houses without central forced air mechanical systems. Stephens et al. and Mueller et al. (1973)
measured O3 deposition loss rates two to four times larger than those measured by Lee et al., but
conducted the measurements while fans were operating which provided mixing and increased
mass transfer. This variation agrees well with measurements conducted by Sabersky et al. (1973)
showing residential O3 deposition loss rates approximately doubled with internal air circulation.
Values of 5.6 hr-1
and 2.8 hr-1
were used in the model for periods with and without ventilation
running, respectively. An O3 penetration rate of 0.72 was used in the model, based on
measurements of eight homes by Stephens et al. Zhao et al. (2007) measured O3 filter removal
efficiencies for eight residential filters and obtained a mean value of 10%, which was used in the
model.
43
6.4 Analysis
6.4.1 Constant Ventilation Rate
In order to evaluate the benefits of the variable ventilation strategy, indoor air quality was
modeled at a constant ventilation rate. The constant ventilation rate was first calculated based on
ANSI/ASHRAE Standard 62.2-2013. Based on a floor area of 186 m2 and three bedrooms, the
design ventilation rate, including infiltration, was 42 L/s (0.32 hr-1
). The standard allows for up
to two thirds of the total ventilation rate to be made up of infiltration. Since an infiltration rate of
5.8 L/s (0.09 hr-1
) was used in the model, the design mechanical ventilation rate was reduced to
30 L/s (0.23 hr-1
).
The concentrations of the three pollutants resulting from the ANSI/ASHRAE 62.2-2013
ventilation rate are shown in Table 2. As this rate, HCHO concentrations exceed the Health
Canada guideline by 12%. To allow the Health Canada guidelines to be used as a minimum
requirement for both the variable and constant ventilation models, the constant ventilation rate
was increased by 17% to 35.3 L/s (0.27 hr-1
). This adjustment brought the steady-state HCHO
concentrations to the Health Canada guideline level of 50 μg/m3.
Table 2. Pollutant concentrations at constant ventilation rates
Constant Ventilation Rate Steady-State Concentration [μg/m
3]
PM2.5 HCHO O3
30.3 L/s (0.23 hr-1
) [ASHRAE 62.2] 17 56 1.2
35.3 L/s (0.27 hr-1
) 16 50 0.7
Health Canada Guideline - 50 40
The low steady-state concentrations of O3 relative to the Health Canada guideline suggest that it
does not have a significant impact on indoor air quality, at least for typical residences located in
Toronto. As a result, HCHO and PM2.5 will be the primary indoor air quality outputs used to
compare ventilation models.
6.4.2 Variable Ventilation Strategy
While the fine wire HRV requires less fan electrical energy and recovers more thermal energy
than most conventional HRVs, further energy savings can be achieved by taking advantage of
DCV in different zones. Based on this assertion, a ventilation control strategy was developed that
44
maximizes energy savings while maintaining pollutant concentrations within the guideline levels
established by Health Canada. As outdoor O3 concentrations in Toronto are too low to pose
indoor air quality concerns, and Health Canada does not publish specific limits for PM2.5, the
variable ventilation strategy focused on maintaining acceptable levels of HCHO. Although not
analyzed in this study, CO2 may also function as a practical basis for designing a DCV strategy.
As reviewed in Section 2.2, there are a number of technologies available for detecting
occupancy. While these technologies currently vary in their reliability, they are expected to
continue to improve over time. For the purposes of this simplified model it is assumed that
occupancy can be accurately detected in each of the four zones. When zones are unoccupied,
maximum energy savings can be achieved by turning off mechanical ventilation. This results in a
buildup of indoor-generated pollutants. In order to meet the Health Canada guidelines,
ventilation must then be increased when the zone is re-occupied. Increased ventilation is most
effective immediately after the zone is re-occupied, as pollutant concentrations are higher. Once
the initial buildup of pollutants has been removed, the constant mechanical ventilation rate
calculated in Section 6.4.1 can be restored. This process is illustrated in Figure 26.
Figure 26. Variable ventilation strategy
The Health Canada short-term guideline recommends one-hour HCHO concentrations be kept
below 123 μg/m3. Since Health Canada does not publish guidelines for exposures shorter than
one hour, an overall maximum HCHO limit of 123 μg/m3 was used, which corresponds to just
over seven hours without mechanical ventilation in the modeled home. After these first seven
45
hours, the minimum mechanical ventilation rate required to keep HCHO concentrations below
123 μg/m3 indefinitely is 0.06 hr
-1. To keep HCHO concentrations from exceeding 123 μg/m
3,
this reduced mechanical ventilation rate of 0.06 hr-1
is recommended after the first seven
unoccupied hours. To simplify operation, the reduced mechanical ventilation rate of 0.06 hr-1
could be applied from the start of the unoccupied period, but would result in slightly lower
energy savings.
To meet the Health Canada long-term guideline, the average concentration of HCHO measured
over the first eight hours after the zone is reoccupied must be below 50 μg/m3. The duration of
increased ventilation required to meet this targets depends on the length of the preceding
unoccupied period as well as the increased ventilation rate. Although heat recovery and fan
efficiency of the fine wire HRV are lower at higher flow rates, more pollutant can be removed
for each unit of ventilation air when concentrations are higher. The result is greater overall
pollutant removal efficiency at higher ventilation rates immediately after the space is reoccupied.
The ventilation rate during this period is limited by the airflow capacity of the fans, the ability of
the space heating system to meet the additional ventilation load and comfort of occupants
exposed to higher air velocities and resulting noise. Further modeling and experimentation is
required to develop an increased ventilation rate that optimizes all these factors. For modeling
purposes, the increased ventilation rate immediately after the zone is reoccupied was arbitrarily
set to 0.82 hr-1
, three times the constant ventilation rate. Based on this ventilation rate, the
required duration of increased ventilation can be determined, as shown in Figure 27.
Figure 27. Required duration of increased mechanical ventilation
0.0
0.5
1.0
1.5
2.0
2.5
0.0 1.0 2.0 3.0 4.0 5.0 6.0 7.0 8.0 9.0 10.0
Re
qu
ire
d D
ura
tio
n o
f In
cre
ase
d
Me
chan
ical
Ve
nti
lati
on
(h
ou
rs)
Duration of Unoccupied Period (hours)
2 hours increased ventilation required for unoccupied
periods greater than 7 hours
46
For unoccupied periods of less than 60 minutes, the increased mechanical ventilation must run
for 40-45% of the time that the space was unoccupied. This ratio decreases with increasingly
long unoccupied periods, resulting in greater energy savings. Because HCHO concentration
buildup is capped at 123 μg/m3 in the unoccupied period, a maximum of two hours of increased
ventilation is required, even for extended unoccupied periods.
6.4.3 Occupancy Schedules
The impact of the decentralized ventilation strategy on annual energy usage is difficult to
quantify without knowing the occupancy pattern of each zone. Without developing a detailed set
of statistically representative residential occupancy scenarios, a conservative estimate of energy
savings can be made by identifying those periods which are likely to be unoccupied for most
homes, most days of the year.
The main living zone was assumed to be occupied from 7AM to 9AM and 6PM to 11PM,
Monday to Friday, and from 7AM to 11PM Saturday and Sunday. The bedrooms were assumed
to be occupied from 9PM to 8AM every day of the week. The schedules are illustrated in Figure
28. The ventilation rate in each zone was modeled based on these schedules and the control
strategy described in Section 6.4.2. The resulting concentrations of PM2.5, HCHO and O3 are
shown in Figure 29 and Figure 30.
Figure 28. Occupancy schedules
Main Living Zone
Bedroom Zones
Occupied
Unoccupied
Main Living Zone
Bedroom Zones
(b) Weekends
(a) Weekdays
4 PM 6 PM 8 PM 10 PM 12 AM
12 AM
12 AM 2 AM 4 AM 6 AM 8 AM 10 AM 12 PM 2 PM
12 PM 2 PM 4 PM 6 PM 8 PM 10 PM12 AM 2 AM 4 AM 6 AM 8 AM 10 AM
47
Fig
ure
29. L
ivin
g r
oom
poll
uta
nt
con
cen
trati
on
s re
sult
ing f
rom
occu
pan
cy s
ched
ule
s
48
Figure 30. Bedroom pollutant concentrations resulting from occupancy schedules
6.5 Results
6.5.1 Energy Consumption
The ventilation heating load and fan electricity consumption were calculated for each ventilation
schedule developed in Section 6.4.3. The results were time- and area-weighted to calculate
whole-house annual energy use for the variable ventilation strategy. For comparison purposes,
the whole-house ventilation heating load and fan electricity consumption was also modeled as a
base case at the constant ventilation rate of 35.3 L/s (0.27 hr-1
). Both heat exchanger technologies
were modeled at the constant ventilation rate to identify the energy savings associated with the
increased sensible heat recovery efficiency and reduced fan electricity consumption of the fine
wire HRV. The results for the three ventilation strategies are shown in Figure 31.
49
(a) Ventilation heat loss (b) Fan electricity consumption
Figure 31. Energy use of constant and variable ventilation strategies
Based on the constant ventilation rate of 35.3 L/s (0.27 hr-1
), the fine wire HRV reduces the
ventilation heat loss and fan electricity consumption by 45% and 43%, respectively.
Implementing the variable ventilation rate strategy described in Section 6.4.3 results in additional
energy savings, reducing both ventilation heat loss and fan electricity consumption by 61% as
compared to a typical conventional HRV operating at the constant ventilation rate.
Di Placido et al. (2014) modeled energy use of a Toronto home built to the R-2000 standard
which is representative of new energy-efficient home construction in Canada. The modeling by
Di Placido et al. indicates an annual space heating intensity of 72 MJ/m2, with 28% of the
heating load coming from ventilation. This is about 30% lower than the ventilation heating load
calculated in this study for a conventional HRV operating at constant ventilation rate. Part of this
discrepancy can be explained by the 17% higher ventilation rate use in this study, as described in
Section 6.4.1. Di Placido et al. also utilized different weather data and included waste heat
produced by fans in the thermal balance. Equation 7, which was used to calculate ventilation heat
loss in Figure 31a, is also limited in that it does not consider interaction between mechanical
systems, building thermal mass or other environmental condition factors such as solar gains.
The analysis done by Di Placido et al. utilized more sophisticated energy modeling software and
the resulting ventilation heating load is considered more accurate. The results of this study are
still useful in that they illustrate the relative benefits of the fine wire HRV and DCV. Since the
model constructed by Di Placido et al. used a conventional HRV with similar performance to the
0
1000
2000
3000
4000
5000
6000
7000
Consant, Convent. HRV
Constant, Fine Wire HRV
Variable, Fine Wire HRV
Ve
nti
lati
on
He
at L
oss
(M
J/ye
ar)
0
100
200
300
400
500
600
700
Consant, Convent. HRV
Constant, Fine Wire HRV
Variable, Fine Wire HRV
Fan
Ele
ctri
city
Use
(kW
h/y
ear
)
-45% -61% -43% -61%
50
one used in the this study’s baseline model, comparable percentile energy savings can likely be
achieved. Applying a 61% reduction in ventilation heating load to the results obtained by Di
Placido et al. indicates that the overall energy required for space heating of an R-2000 home
could be reduced by 17%. Thus, using the fine wire HRV in combination with DCV can be
expected to reduce annual heating fuel consumption by 12.3 MJ/m2.
Statistics Canada (2011) lists the average Ontario household’s annual electricity consumption as
8333 kWh. Although this is an average across houses of different sizes and ages, it provides a
rough approximation of electricity consumption in the house modeled in this study. Assuming an
annual household electricity consumption of 8333 kWh, the fine wire HRV and DCV have the
potential to reduce total household electricity consumption by approximately 5%. This
percentage may be higher for energy efficient housing with lower overall electricity consumption
due to high efficiency appliances and lighting.
6.5.2 Indoor Air Quality
Pollutant concentrations resulting from the constant and variable ventilation strategies were also
compared. Concentrations were averaged over the occupied periods and are shown in Figure 32.
(a) PM2.5 (b) HCHO (b) O3
Figure 32. Pollutant concentrations of constant and variable ventilation strategies
The average concentration of PM2.5 during the occupied periods is reduced by 21% in the
variable ventilation strategy. As PM2.5 is suspected of having adverse health effects even at low
concentrations (Health Canada), this reduction is significant in improvement in occupant health.
0
2
4
6
8
10
12
14
16
18
Constant Ventilation
Variable Ventilation
PM
2.5
(μg/
m3)
0
10
20
30
40
50
60
Constant Ventilation
Variable Ventilation
HC
HO
(μ
g/m
3)
0.0
0.5
1.0
1.5
2.0
2.5
Constant Ventilation
Variable Ventilation
O3
(μg/
m3)
-21% +41% +3%
51
There is a 3% increase in average HCHO concentrations during occupied periods in the variable
ventilation strategy. This increase results from the short occupancy periods in the main living
space during the weekday mornings and evenings. HCHO concentration builds up during the
unoccupied period and results in high initial HCHO levels when the space is first occupied. Since
these occupied periods are short, the high initial concentration results in higher average
concentrations, even though HCHO concentration is reduced quickly over the first two hours.
Since these are short events by definition, they are unlikely to contribute significantly to an
individual’s overall HCHO exposure.
Although there is a 41% increase in average O3 concentration during occupied periods due to an
increase in the average ventilation rate, the concentration is still more than an order of magnitude
less than the Health Canada (2014) guideline. Given the very low O3 concentration in both
strategies, the increase is not expected to have a significant impact on health.
6.6 Limitations
There are a number of limitations to the modeling approach. The parameters used to model
indoor air quality are mean values, and do not capture the distributions of values. Furthermore,
parameters are assumed constant over time which may not capture important relationships
between different parameters. For example, indoor HCHO emission rates likely decrease with
increasing concentration. Similarly, indoor PM2.5 emissions are likely higher during occupied
periods due to activities such as cooking or cleaning which generate particles. In both of these
examples the assumptions are conservative in that they underestimate the benefits of the
decentralized ventilation system, but there may be other less understood relationships with
unknown implications.
Because the proposed variable ventilation strategy is based on the mean HCHO emission rate
used in the model, it is not optimized for homes with different HCHO emission levels. In cases
where emissions are lower, the proposed strategy may result in wasted energy, while in cases
where emissions are higher it may result in HCHO concentrations exceeding the Health Canada
(2014) guidelines. This challenge is also faced by constant ventilation rate guidelines and is
largely unavoidable. Sensitivity analysis of the relative impact of HCHO emission rates on the
optimized ventilation strategy would help quantify the significance of this concern.
52
The outdoor pollutant concentrations used in the model are based on average conditions in
Toronto. While Toronto has relatively low average ambient levels of O3 and PM2.5, outdoor
concentrations of these pollutants can be higher during particular times of the day. For example,
traffic-generated outdoor pollutant levels may peak during the morning and afternoon rush hours.
If the afternoon peak coincides with occupants returning home and increased ventilation rates, it
may introduce higher levels of these traffic-generated pollutants into the indoor environment.
Similarly, for those locations in the world where outdoor pollution levels are very high, the
proposed strategy of increasing the ventilation rate when a zone is initially occupied may
actually expose occupants to worse indoor air quality. Care should be taken to consider
differences in outdoor air quality when extrapolating these results to other locations.
The calculations of ventilation heat loss are based on a simplified equation which does not
consider the building systems in detail or the effects of other environmental factors such as solar
gains. More accurate results could be obtained by performing a whole-building energy model.
The occupancy schedules assumed in the model also limit the relevance of the results, which
may differ under different occupancy scenarios.
The many limitations and unknown variables inherent to the modeling approach mean that there
is some uncertainty associated with the results. Although the results may not be representative of
all installations of the fine wire HRV, they do illustrate the potential for significant reductions in
ventilation heating load and fan electricity consumption. Further modeling and field testing is
necessary to provide more reliable and representative results.
7 Discussion
The results of this study indicate that the fine wire HRV, in combination with a variable and
demand controlled ventilation strategy, may offer significant benefits over conventional
ventilation systems. The apparent effectiveness of the technology raises the question of why it
has not been successfully commercialized yet. Although the fine wire heat exchanger was first
patented in North America over 15 years ago, it is not yet commercially available and to the
author’s knowledge has received limited industry attention. This may be in part a reality of trying
to manufacture and promote an unconventional product with limited resources. That being said,
it is also important to consider some of the challenges which may impact the technology’s uptake
in the residential market.
53
Unlike a conventional ventilation system, the fine wire HRV is designed to be located in close
proximity to building occupants. Similar to a window air conditioner or portable air cleaner, the
noise generated by fans and air movement may be disruptive, particularly in bedrooms. This
issue is likely most significant in Northern Europe, where people are generally less tolerant of
noise, as evidenced by the importance of acoustics in European ventilation standards (eg.
Passivhaus Institut 2009). Even in North America, some building owners, particularly those
interested in high-end energy efficient housing, may find the fine wire HRV less attractive as a
result. Further investigation should be conducted to quantify noise levels and explore options for
dampening sound or locating the HRV to minimize acoustical impacts.
The cost of installation and maintenance is another issue which may dissuade home owners from
installing the fine wire HRV. Although the final retail price of the unit itself should be
comparable to conventional HRVs, a greater number of units are required to take advantage of
decentralized ventilation, increasing the installation cost. This increased cost will be partly offset
by reduced ductwork requirements. Maintenance costs will also be higher, as there are a larger
number of fans, electrical components and filters to be serviced.
The challenge of providing multiple HRVs is further compounded by the need for condensate
drainage in cold climates. Moisture is produced in buildings by activities such as showering and
cooking, as well as by the biological processes of occupants themselves. The result is relatively
warm and moist indoor air, regardless of outdoor humidity levels. When outdoor temperatures
are low, HRVs function by extracting energy from the warm exhaust air to heat the cool supply
air. If the exhaust air is cooled below its dew point it will not be able to contain the water vapour,
resulting in condensation which must be drained away. Providing drains to multiple HRVs
located on exterior walls throughout the home may be costly, particularly in a retrofit situation.
The fine wire heat exchanger is only capable of recovering sensible heat, making it most relevant
for heating dominated climates. Given the fundamental heat recovery concept, an option for
latent heat recovery does not appear feasible. Although the layers separating adjacent airstreams
could potentially be made vapour permeable, the very small surface area would likely still make
significant latent heat transfer difficult. Thus, the technology has a smaller potential market than
rotary and plate heat exchangers which can be designed to recover both sensible and latent heat.
54
8 Conclusions
The fine wire HRV represents a novel concept for decentralized ventilation with heat recovery.
The ventilator uses copper wires to transfer sensible energy between the supply and exhaust
airstreams. By integrating the fine wire HRV into the façade to create a “breathing wall”, the
extensive ductwork and associated space requirements of conventional ventilation systems can
be avoided. This also allows each room or zone of the home to have its own ventilator,
facilitating demand controlled ventilation. This initial scientific review of the fine wire HRV
evaluated its performance in several areas as well as its suitability for decentralized and demand
controlled ventilation heat recovery in residential buildings.
Performance testing was conducted in an indoor laboratory. Fan electricity consumption using
brushless DC motors was found to be 1.1 W per L/s for airflow rates below 30 L/s, putting the
fine wire HRV within the top 5% of North American HRVs with respect to fan energy. The fine
wire HRV was also found to have a sensible heat recovery efficiency as high as 82% for airflow
rates of 7-10 L/s. At this performance level, it is within the top 2.5% of North American HRVs
with respect to heat recovery. Airstream cross-leakage was found to be 0.5-1.5% and not to have
a significant impact on performance.
A model of ventilation energy usage and indoor air quality was developed and used to compare a
conventional HRV to the fine wire HRV. The model results indicate that the fine wire HRV in
conjunction with demand controlled ventilation could reduce the ventilation heating load by 61%
and the total heating load by 17%. Fan electricity consumption was also reduced by 61%,
corresponding to a 5% reduction in total household electricity use. Indoor air quality was also
improved, with a 21% reduction in average PM2.5 concentrations.
The technology has yet to be successfully commercialized and there are a number of challenges
which may slow adoption of the fine wire HRV. Due to the proximity to building occupants,
noise generated by fans and air movement may be disruptive, particularly in bedrooms. The cost
of installation and maintenance of multiple ventilation units is another issue which may dissuade
home owners. Finally, the inability to recover latent heat means that the fine wire HRV is only
appropriate for heating dominated climates.
55
9 Future Research
9.1 Condensation and Freezing
One issue that has not been investigated in this study is the fine wire HRV’s capacity to handle
condensation and freezing. During moderate winter weather, water vapour will condense on the
heat exchanger surface and can be drained away. High efficiency heat exchangers and cold
outdoor temperatures mean exhaust air may be cooled below 0 ⁰C however, resulting in freezing
and an accumulation of ice on the heat exchanger surface. This ice eventually obstructs airflow
and results in fouling of the HRV. Different HRV designs address ice accumulation in different
ways. In North America, the most common approach is to preheat the supply air inlet above a
minimum temperature so that freezing does not occur in the exhaust airstream. To simplify
installation, this preheating is often done with electric resistance heat, making it a costly process
relative to heating with natural gas. Preheating the supply air inlet also reduce the total heat that
can be recovered from the exhaust airstream, making the HRV less efficient. Other HRV designs
either stop or reduce supply airflow temporarily, increasing the temperature of the exhaust
airstream and allowing the ice to melt. This must be done periodically as ice re-accumulates,
reducing the HRVs overall heat recovery efficiency in cold weather.
The impact of condensation and freezing of water vapour on the fine wire heat exchanger has not
been explored in this study. The developer has suggested that the very rapid cooling of the
exhaust air as it passes through the 16 mm thick heat exchanger results in air exiting in a super-
saturated state. If this is true, it could mean reduced condensation and ice accumulation and
improved performance in cold weather. Tests should be conducted to determine under what
conditions ice begins to accumulate, and how much supply airflow must be reduced to defrost
the heat exchanger. These results can then inform an effective control strategy and an energy
modeling effort which accounts for the reduction in annual efficiency due to freezing.
56
9.2 Field Installation
The model of whole-building energy use and indoor air quality provides an indication of the
potential of the fine wire HRV. As with any modeling exercise, increased confidence in the
results requires field validation. Ideally this would be done by installing a number of fine wire
HRVs in a typical new home which also had a conventional HRV system and associated
ductwork. The house could then be operated intermittently with the two systems, and the effects
on energy consumption and indoor air quality could be measured.
Undoubtedly many of the assumptions made in the model will have to be refined. The
performance measurements taken in the lab may also have to be adjusted, as technology often
behaves differently in the field. For example, all the performance measurements taken as part of
this study were conducted under steady-state conditions. It is unclear what effect continuously
changing airflows in the field might have on performance. Occupancy may also be a challenge to
accurately detect, and some level of control failure may have to be incorporated into the model.
57
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62
Appendix A:
Photos of Prototype, Testing and Installations
63
Manifold (back side) Manifold (front side)
Heat exchanger core without manifold Heat exchanger core detail
64
Fans mounted back-to-back Plywood sandwich layer with wiring
Installed fan Fan power supplies and controls
Location of static pressure probe Test setup with calibrated fan
65
Completed prototype Prototype with thermocouples
Example installation (by others) Example installation (by others)
66
Appendix B:
Sensible Heat Recovery Efficiency Test Data
67
Test 1 – Airflow rate: 12 cfm (5.7 L/s) – Heat at exhaust inlet: 200 W
Test 2 – Airflow rate: 15 cfm (7.1 L/s) – Heat at exhaust inlet: 215 W
20
22
24
26
28
30
32
34
36
38
40
0 50 100 150 200 250 300 350 400 450
Me
asu
red
Te
mp
era
ture
(°C)
Time (seconds)
t1(a)
t1(b)
t1(c)
t2(a)
t2(b)
t2(c)
t3(a)
t3(b)
t3(c)
t4(a)
t4(b)
t4(c)
t3
t2
t4
t1
20
22
24
26
28
30
32
34
36
38
40
0 50 100 150 200 250 300 350 400 450
Me
asu
red
Te
mp
era
ture
(°C)
Time (seconds)
t1(a)
t1(b)
t1(c)
t2(a)
t2(b)
t2(c)
t3(a)
t3(b)
t3(c)
t4(a)
t4(b)
t4(c)
t3
t2
t4
t1
68
Test 3 – Airflow rate: 20.5 cfm (9.7 L/s) – Heat at exhaust inlet: 260 W
Test 4 – Airflow rate: 26 cfm (12.3 L/s) – Heat at exhaust inlet: 300 W
20
22
24
26
28
30
32
34
36
38
40
0 50 100 150 200 250 300 350
Me
asu
red
Te
mp
era
ture
(°C)
Time (seconds)
t1(a)
t1(b)
t1(c)
t2(a)
t2(b)
t2(c)
t3(a)
t3(b)
t3(c)
t4(a)
t4(b)
t4(c)
t3
t2
t4
t1
20
22
24
26
28
30
32
34
36
38
40
0 100 200 300 400 500
Me
asu
red
Te
mp
era
ture
(°C)
Time (seconds)
t1(a)
t1(b)
t1(c)
t2(a)
t2(b)
t2(c)
t3(a)
t3(b)
t3(c)
t4(a)
t4(b)
t4(c)
t3
t2
t4
t1
69
Test 5 – Airflow rate: 28.5 cfm (13.5 L/s) – Heat at exhaust inlet: 300 W
Test 6 – Airflow rate: 33 cfm (15.6 L/s) – Heat at exhaust inlet: 320 W
20
22
24
26
28
30
32
34
36
38
40
0 50 100 150 200 250 300 350 400
Me
asu
red
Te
mp
era
ture
(°C)
Time (seconds)
t1(a)
t1(b)
t1(c)
t2(a)
t2(b)
t2(c)
t3(a)
t3(b)
t3(c)
t4(a)
t4(b)
t4(c)
t3
t2
t4
t1
20
22
24
26
28
30
32
34
36
38
40
0 50 100 150 200 250
Me
asu
red
Te
mp
era
ture
(°C)
Time (seconds)
t1(a)
t1(b)
t1(c)
t2(a)
t2(b)
t2(c)
t3(a)
t3(b)
t3(c)
t4(a)
t4(b)
t4(c)
t3
t2
t4
t1
70
Test 7 – Airflow rate: 39 cfm (18.4 L/s) – Heat at exhaust inlet: 345 W
Test 8 – Airflow rate: 48 cfm (22.7 L/s) – Heat at exhaust inlet: 370 W
20
22
24
26
28
30
32
34
36
38
40
0 50 100 150 200 250
Me
asu
red
Te
mp
era
ture
(°C)
Time (seconds)
t1(a)
t1(b)
t1(c)
t2(a)
t2(b)
t2(c)
t3(a)
t3(b)
t3(c)
t4(a)
t4(b)
t4(c)
t3
t2
t4
t1
20
22
24
26
28
30
32
34
36
38
40
0 50 100 150 200 250 300 350 400
Me
asu
red
Te
mp
era
ture
(°C)
Time (seconds)
t1(a)
t1(b)
t1(c)
t2(a)
t2(b)
t2(c)
t3(a)
t3(b)
t3(c)
t4(a)
t4(b)
t4(c)
t3
t2
t4
t1
71
Test 9 – Airflow rate: 56 cfm (26.4 L/s) – Heat at exhaust inlet: 370 W
20
22
24
26
28
30
32
34
36
38
40
0 50 100 150 200 250 300 350 400 450
Me
asu
red
Te
mp
era
ture
(°C)
Time (seconds)
t1(a)
t1(b)
t1(c)
t2(a)
t2(b)
t2(c)
t3(a)
t3(b)
t3(c)
t4(a)
t4(b)
t4(c)
t3
t2
t4
t1