REFERENC1

NBS TECHNICAL NOTE 710-5

U.S. 'ARTMENT

OF

OMMERCE

100

U5753

no- 710-^

BUILDING

RESEARCH TRANSLATION

Use of an

Air-to-Air Heat Exchanger

To Recover Heat from Air

Exhausted by Mechanical

Ventilation
NATIONAL BUREAU OF STANDARDS

The National Bureau of Standards1 was established by an act of Congress March 3, 1901. The Bureau's overall goal is to strengthen and advance the Nation's science and

technology and facilitate their effective application for public benefit. To this end, the Bureau conducts research and provides: (1) a basis for the Nation's physical measure- ment system, (2) scientific and technological services for industry and government, (3) a technical basis for equity in trade, and (4) technical services to promote public safety. The Bureau consists of the Institute for Basic Standards, the Institute for Materials Research, the Institute for Applied Technology, the Center for Computer Sciences and

Technology, and the Office for Information Programs.

THE INSTITUTE FOR BASIC STANDARDS provides the central basis within the United States of a complete and consistent system of physical measurement; coordinates

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provides liaison between NBS and national and international engineering standards

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THE CENTER FOR COMPUTER SCIENCES AND TECHNOLOGY conducts re- search and provides technical services designed to aid Government agencies in improving cost effectiveness in the conduct of their programs through the selection, acquisition, and effective utilization of automatic data processing equipment; and serves as the principal focus within the executive branch for the development of Federal standards for automatic data processing equipment, techniques, and computer languages. The Center

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THE OFFICE FOR INFORMATION PROGRAMS promotes optimum dissemination and accessibility of scientific information generated within NBS and other agencies of

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1 Headquarters and Laboratories at Gaithersburg, Maryland, unless otherwise noted; mailing address Washington, D.C. 20234.

~ Part of the Center for Radiation Research.

3 Located at Boulder, Colorado 80302.
\0D

in

BUILDING RESEARCH TRANSLATION

Use of an Air-to-Air Heat Exchanger

To Recover Heat from Air

Exhausted by Mechanical Ventilation

P. Garrivier

Centre Scientifique et Technique du Batiment

Paris, France

Translated by the Joint Publication Research Service, U.S. Department of Commerce,

Washington, D.C. 20234

SrATti O* '

U.S. DEPARTMENT OF COMMERCE, Peter G. Peterson, Secretary

U-'S. NATIONAL BUREAU OF STANDARDS, Lawrence M. Kushner, Acting Director,

Issued July 1972
National Bureau of Standards Technical Note 710-5

Nat. Bur. Stand. (U.S.), Tech. Note 710-5, 62 pages (July 1972) CODEN: NBTNAE

For Bale by the Superintendent of Documents, U.S. Government Printing Office, Washington, D.C. 20402 (Order by SD Catalog No. C13.46:710-5). Price 60cents.
FOREWORD

The United States /French Cooperative Program on Building

Technology entails an exchange of personnel between the

National Bureau of Standards (Building Research Division)

and the Centre Scientifique et Technique du Batiment

(CSTB) of France. The program also involves the exchange

of information between the two research organizations.

It is felt that some of the documented information can

be usefully shared with the U.S. building industry; and,

therefore, certain papers were selected for reproduction

in media on sale to the public by the Government Printing

Office. It should be understood that the CSTB documents

made public through such media as this TECHNICAL NOTE

do not necessarily represent the views of the National

Bureau of Standards on either policy or technical levels.

At the same time, building researchers at the National

Bureau of Standards consider it a public service to share

with the U.S. building industry certain insights into

French building technology.

<JAMES R. WRIGHT

Chief, Building Research Division

Institute for Applied Technology

National Bureau of Standards

III

BUILDING RESEARCH TRANSLATION

USE OF AN AIR-TO-AIR HEAT EXCHANGER

TO RECOVER HEAT FROM AIR EXHAUSTED

BY MECHANICAL VENTILATION-'

by

P. Garrivier, Engineer

When a 'building is equipped with a mechanical system which

both exhausts and supplies outdoor air, it may be profitable to

install a heat exchanger between the two circuits. The CSTB

has perfected such an exchanger, which meets the two principal

requirements of the problem: good efficiencies and elimination of the polluted air intake. A method of calculating the

optimal dimensions of the exchanger has also been devised.

Key words: Exchanger dimensions; frost formation; heat

exchanger; heating costs; mechanical systems; pollution;

pressure loss; ventilation.

—This paper is translated from the French original and

is published under the Building Research Division, Centre

Scientifique et Technique du Batiment information exchange

program.
1. INTRODUCTION: FEASIBILITY OF EXCHANGER LINKED TO DEVELOPMENT OF

MECHANICAL VENTILATION .

Heat loss through air change represents, in general, some 30 per

cent of the overall heat loss in a multiple-dwelling huilding and

approximately 20 per cent of such loss in a single house.

Recovery of heat from the outgoing air by means of a heat exchanger

installed between the exhaust air and the incoming air can therefore

achieve significant savings. With an exchanger efficiency of TO per

cent —the actual efficiency attained by the CSTB-developed exchanger—the

resultant saving in heating is of the order of 21 per cent (0.30 x 0.70)

in a multiple-occupancy building, and lk per cent (0.20 x 0.70) in a

single house.

With the improved insulation of walls, floors and ceilings and the

increased demands for air purity, these figures should increase

appreciably in years to come.

The cost of the exchanger per se—as we shall see later—is not high.

But its utilization does assume that the building is equipped with a

mechanical forced (blown) air intake and a mechanical exhaust. Use of

the latter is expanding rapidly in France. In 1971, 15 to 20 per cent of

the residential construction will have such equipment. In 10 years the

figure will perhaps be 30 to 1+0 percent, with the figure in multipledwelling buildings probably rising to over 50 per cent. On the other hand,

the forced air-delivery system, i.e. by means of a mechanical ventilation,

is currently nonexistent, so to speak. The reason for this is its high

cost, which is not the case of mechanical extraction or exhaust equipment.

2
In fact, exhausting air by natural circulation (draft) requires vertical

ducts or conduits , which in a multiple-dwelling "building cost nearly as

much as a mechanical exhaust system. On the other hand, the delivery of

air is generally accomplished by means of simple openings on the outside

walls. Their cost is much lower than that of a mechanical forced-air

delivery system. We do believe, however, that mechanical forced-air

delivery will experience some expansion in the coming years for the

following reasons

:

1. A first reason is the need to shield oneself against noise in

areas such as those close to airports, superhighways, and other roads and

highways. To do this one must satisfy air-purity and thermal-comfort

requirements without opening any windows . Thus a mechanical forced-air

supply is almost indispensable, first, in order to eliminate ventilation

openings in outside walls , then to obtain greater flexibility in heating

(that is so as not to be obliged to control heating by opening windows),

and lastly, to obtain the large number of necessary air changes in the

summertime (which are usually obtained by opening windows).

2. A second reason is the requirement to humidify the air in cold

dry weather. This need will become increasingly greater because, first

of all, comfort demands are growing, and then because the higher temperatures desired by occupants lower relative humidity, and lastly because

while good ventilation (i.e. mechanical exhaust in kitchens, bathrooms,

toilets) does eliminate odors, it also lowers the humidity in the main

rooms (kitchen, bathroom, etc.)

The calculations found in the appendix to this document, show that

during an average winter in Paris and within a room heated to 21°C in a

3
properly ventilated dwelling, air humidity is below 20 per cent during

eight days and ranges from 20 to 25 per cent for l6 days, from 25 to 30

per cent for 28 days, and from 30 to 35 percent for 30 days.

If 35 per cent is deemed to he the minimum desirable humidity, then

the air must be humidified during 9^+ days, and within the room under

consideration the following amounts of water must be evaporated daily:

1.6 liters when humidity is 20 per cent, 1.1 liters when it is 25 per

cent, and 0.5 liter when it is 30 per cent.

The number of days in which humidification is necessary and the

quant it it ies of water that must be evaporated, are both considerable. It

is difficult to do this properly without having air mechanically introduced

into the dwelling.

3. A third reason is the desirability of filtering the air, a

requirement that will grow as pollution increases.

k. A fourth reason is the need to control ventilation in very windy

sites, e.g. in high-rise buildings, so as to avoid disagreeable drafts

and recurrent temporary drops of the indoor temperature.

5. A fifth reason is the attempt to find an exceptionally flexible

means of heating, at least of secondary or supplementary heating.

Several of these reasons take on increased importance when electric

heating is involved. In fact a higher degree of comfort has to be

provided to warrant what is generally a highly expensive proposition.

Furthermore, to limit this expenditure one must have control over all

heat loss through air changes. And lastly, because of the advances in wall,

ceiling, and floor insulation, any heat loss through air changes is

relatively greater, hence the added value of the exchanger and
consequently of a forced air intake. Up to now we have talked only about

the residential aspects of this question. But in other areas of activity

the development of forced-air systems may be more rapid: in hospitals

for comfort and hygienic reason; in schools —frequently lacking thermal

inertia and very densely occupied—for reasons of thermal comfort in the

summer and also for reasons of flexibility in heating; and in offices

that are being air-conditioned with increasing frequency and where, as a

result, a forced-air intake is available almost automatically.

And finally, we wish to add that the heat exchanger may in certain

types of construction, or with certain types of equipment, be used for

summer comfort, with or without air conditioning. But this subject is

still under study and we shall not elaborate on it in this document.

2. VARIOUS HEAT-RECOVERY SYSTEMS DESERVING CONSIDERATION: SYSTEMS

CURRENTLY AVAILABLE FOR DWELLING USE .

We shall survey existing exchanger systems by examining advantages

and disadvantages of each one from various points of view, particularly: —that of efficiency which is obviously of capital importance; —that of pressure loss which increases ventilation costs

and makes acoustic problems more difficult to resolve; —that of frost formation which affects efficiency; —that of sanitation: the forced-air intake must not be

polluted by the exhausted air.
2.1 Solutions Based on the Thermal Inertia of a Material .

There are many such solutions. They may involve having the incoming

air and the outgoing air alternately pass over a divided mass so that the

warm stream yields its heat to the mass and the cold stream is warmed

upon contact with that mass. Different arrangements are possible. For

example, one apparatus consists of two chambers containing the material

and through which each stream passes for a certain period of time with

periodic reversal (such solutions are utilized industrially, especially

in the iron and steel industry)

The efficiency, small pressure-loss, and relatively meager space

requirement of these solutions, are all advantages. But the solutions do

have the disadvantage of recirculating odors and impurities as the new

air passes over the material polluted by the exhausted air. Moreover,

recirculation of a certain volume of gas occurs at the moment of aircurrent inversion or reversal. This disadvantage can be limited by

purging the exchanger at the moment of reversal, but this requires

additional apparatus. In addition, the exchange mass of material may

gradually become fouled and clogged. This situation can be remedied by

either the use of a washing device or by simply replacing the

material, both solutions which do have their disadvantages. Figure 1 has

an illustrative diagram.

Another application of this same principle is the rotary-exchanger

solution. It consists of a porous wheel which rotates about an axis

paralled to two gaseous streams separated by a wall that diametrically

intersects the wheel (see Figure 2). In this solution, the air streams
follow a fixed course and it is the collector (accumulating) material that

passes from one stream to the other at variable speed. This solution

which, under good conditions, is capable of achieving 80 per cent

efficiency, has the same disadvantages as the previous solution. A

purger permits limiting the recirculation of stale air at the point where

the wheel crosses the boundary between the stream of exhausted air and the

stream of new air. But there is still a risk of recirculation on a level

with the joint or seal between the rotating part and the barrier separating

the two streams. This risk may be limited if the pressure of the new air

exceeds that of the exhausted air (which is generally the case).

These exchangers are widely used in the Scandinavian countries (a

Swedish firm makes them) , and in North America (there is a Canadian

manufacturer). A study by the National Research Council of Canada's

Division of Building Research revealed that even in the best of cases

there was still a two to four per cent recirculation of the bacteria

contained in the exhausted air. This situation indicates that highefficiency filters should be added to the system. Moreover, the fouling

and clogging of the air passages that pass through the wheel (with a

resultant increase in pressure loss through the wheel plus transmission

of odors) can be prevented only by prior filtering of the exhausted air

(this is an important constraint). In practice this solution is not

used in kitchens, a fact that greatly reduces its value.

Frost is apt to form within these types of exchangers when the moist

outgoing air passes over that part of the exchange material whose temperature is close to the outside temperature which is below 0°C. This
condition can "be corrected by use of a preheating system or "by shunting

the system below a certain outside temperature. Unfortunately these

remedial steps decrease exchanger efficiency.

2.2 Systems Employing an Intermediary Heat-Carrying Fluid .

This solution consists in having a heat-carrying fluid circulate

between two finned tube coils, one set within the warm-air stream, the

other within the cold air stream. A pump serves to circulate this fluid.

(See Figure 3.

This system offers significant advantages over the previous one:

—the exhaust ducts and forced-air intake ducts in which heat

exchange occurs may De placed at some distance from each

other, —recirculation of odors and contaminants is impossible.

On the other hand, this solution is not very economical. Initial

investment costs are high and the conventional finned tube coils cause

considerable pressure loss whenever they are required to work effectively.

Frost problems and fouling difficulties are resolved as in the first

solution.

2.3 Systems Using Direct Heat Transfer Through a Conductive Wall Surface .

This is the family of "plate" heat exchangers. They exist in

various forms and with counterflow, parallel flow, crossflow. . . There

are manifold versions of these exchangers, mainly with applications

rather different and remote from those being considered here (we are aware

of no applications of this type of exchanger in the building field).

8
One variation consists in having the exhaust and forced-air intake

ducts arranged coaxially so that transfer occurs throughout the building's

height, with the gases flowing in opposite directions, counterflow. We

did not consider this version worthy of more detailed study.

3. DEVELOPMENT OF A PROTOTYPE

3.1 Choices

Selection of the Type of Exchanger

This survey of the different types of heat exchangers revealed that: —exchanger employing the inertia of a material can have a high

degree of effectiveness but have disadvantages from the

hygienic and odor point of view; —exchanger with an intermediary heat-transporting fluid are

not- economically satisfactory in our application.

That leaves the plate-type exchangers whose efficiency and economy

depend essentially upon the nature and geometry of the plates. We decided

to especially study this last solution in greater detail.

Selection of a Configuration

There are three basic configurations open to consideration:

— counterflow:

— parallel Flow:

crossflow:
There are other solutions that combine these three single solutions.

For reasons of manufacturing simplicity we did not retain them for

further consideration.

The parallel-flow exchanger is not conducive to an acceptable

efficiency (for example, in the case of two identical flow rates,

efficiency is limited to 50 per cent).

The crossflow exchangers when compared with the counterflow exchanger

has the advantage of an easy arrangement of apertures , with solely one

gaseous flow issuing onto each surface. This is a solution frequently

adopted for gas-to-gas exchangers. We chose this solution as the one to

use in building a research model.

Note: from results obtained with this model, it was subsequently

possible to deduce those results obtained in different configurations

with plates having the same characteristics.

Selection of a Material

Factors that must be considered: —cost of the material, —its ease of employment, especially the ease with which it

can be formed or molded, —its mechanical behavior in the planned field of utilization, —its chemical behavior, —its conductivity.

It should be noted that if small thicknesses of material are used

to make the plates, then the last factor is only of secondary importance:

the thermal resistances encountered are mainly those of heat transfers by

convection between air and plates.

10
The cost of the material and its durability must be considered

together: —either the material used is expensive (per se or in its

application! and designed to last several years

without deterioration, —or else an inexpensive material is selected and it is conceded

that the part of the exchanger consisting of the pile of

plates will he replaced at rather close time intervals (every

year , for example )

.

It must he noted that in the case of the "long-lasting" plates , the

exchanger cleaning problem has to be resolved. If filters are not

installed, then a capability must be provided for periodically cleaning

of the entire plate assembly. This cleaning operation must be taken into

account when determining the economic feasibility of such plates, and

moreover, it imposes certain constraints in the manufacture of the

exchanger. By changing the plates, e. g. annually, such constraints and

cleaning operations are avoided.

Consequently, the "expendable" plates solution was the one chosen.

We directed our effort toward finding a very inexpensive material

that was also simple to produce commercially, namely a plastic material

that could be "hot-formed" or molded. Tests were conducted on small-size

samples to determine their plasticity (capability to be shaped or molded),

and their mechanical behavior when not very thick. These tests led us to

choose—not exclusively—0.17 mm-thick sheets of cellulose acetate. (This

product is available commercially under the trade name of Rhodialine.

It has a density of 1.29 x 10-3 kg/m^ and a maximum utlization temperature

of 60°C. ) 11
Selection of a Plate Geometry

The following requirements had to be taken into account: —plates had to he easily obtainable by a "hot-forming" process

and fabrication of the mold must not be too complex; —there must be maximum convection but all pressure loss or

drop nevertheless had to be quite low; —plates had to be "self-supporting," in other words the use of

braces or ties must not be required to join these plates

together, and they must be capable of being piled or stacked

naturally.

We adopted a geometry reminiscent of an egg carton. It was obtained

by the simultaneous and opposing compression of two matrixes having

identical networks (grids) of studs with both networks being staggered a

half module (the patent statement annexed hereto gives a more detailed

description of this type of surface).

3.2 Building of the Prototype

The model exchanger was fabricated in the CSTB, using a variety of

in-house equipment and machine tools. The limitations of these tools and

equipment did impose the dimensions adopted for the plates. These

dimensions were, however, of a highly suitable order of magnitude as

subsequently demonstrated by the study of optimum exchanger dimensions.

Fabrication of a Matrix

Use of the CSTB's available "hot-forming" equipment necessitated

separate casting of each plate. In large-scale industrial production this

process would be advantageously replaced by a continuous calendering of the

plates.

12
The matrix was made in several stages:

1. Fabrication of a model shaped like the plates. To accomplish

this, 931 cylindrical plexiglass studs or pins, 10 mm in diameter and

8 mm high, were glued to a wooden base (19 mm lathing) in a 30 mm-module

grid pattern. A 3 mm-thick sheet of butyl rubber was spread over the

studs, and then with the help of a template, a tack was nailed into the

center of each square formed by four adjacent studs, these fastening the

butyl rubber down against the wood at that point ( see Figure 5 )

•

2. A matrix was then cast upon this form. A metal (U-bar) frame

was placed along the perimeter of the model (see Figure 6) and a Uo kg

mixture of polyester resin and aluminosilicate with a polymerizing agent

and accelerator, were poured inside the metal frame. One hour later

polymerization was complete and the polyester matrix could then be

separated from the form.

3. To make the matrix suitable for use on CSTB equipment, it was then

necessary to pierce a small-diameter hole in the bottom of each concave

part of the matrix (this hole allows a reduced pressure to be created in

each of the small cavities of the matrix during "hot-forming").

Hot-Forming of the Plates

This operation was carried out on the CSTB's "Illig" hot-forming

machine. Matrix dimensions had been selected so that available equipment

could be used (see Figure 7).

The operation consisted of the following successive steps: —infrared heating at 110°C ( on movable carriage )

;

—slight distention of the sheet to be formed; —suction from under the matrix,

13
—cooling period.

Plates were then cut into 915 by 915 mm square units.

Assembling the Plates

The useful part of the prototype exchanger consists of a pile of 22

plates. These were arranged in this pile by means of four corner guides

positioned perpendicular to the plates. These guides are ho mm X h() mm pq.

cross sections matching the small areas cut out at the corners of each

plate. Two successive plates were symmetrical with reference to the

plane containing their points of contact, points which are the high points

("summits") of the surface. Some of these points were roughly glued to

each other.

The parallel edges of successive sheets were glued together in pairs,

the first in one direction, the second in another, so as to enable

gaseous streams to flow from one layer to another in perpendicular

directions.

A rubber "leading edge" was mounted on the adhesive-bonded edges to

provide a better mechanical hold and a good seal. This precautionary

step could be dispensed with in large-scale production where suitable

binding equipment would be used.

After the plates had thus been assembled, the unit was set into a

frame or case which provided rigidity to the entire assembly. This case

could be easily removed: a threaded stem or bolt mounted on each corner

guide permitted tightening the frame with a wing nut. The top and bottom

plates were insulated from the frame by a 3 cm-thick sheet of expanded

polystyrene.

14
A foam-rubber seam served to seal the corners where plates and

bracing guides joined.

Figures 8 and 9 depict this assembly method.

Exchanger Connections

The parallelepiped formed by the plates is finished off with sheetmetal sleeves connecting the circular cross sections of the ventilation

ducts to the rectangular inlet cross sections of the exchanger (see

Figure 10).

k. TESTS AND MEASUREMENTS ON THE PROTOTYPE BUILT AT THE CSTB

k.l Test Equipment (Figure 11

)

Testing facilities reproduced a central forced-air intake and

exhaust system for a building or dwelling. It therefore included a forcedair intake circuit with its own fan and an exhaust circuit with its fan.

Dampers were used to regulate the flow rate in each of the circuits.

Tests reproduced normal application conditions. An electric heater

was installed in the exhaust circuit. Since its heat output could be

continuously controlled, it was thus possible to reproduce inside conditions

as desired.

Afterward a system for reproducing variable humidity conditions

(greater than those of the outside air) was added to this air-heating

system. This humidity system consisted of an insulated tank, equipped

with a variable-power electric resistor. This device produced a known

quantity of water vapor which was injected by suction into the exhaust

circuit downstream from the electrical resistance heaters (see Figure 12).

15
Since a condensation process might occur within the exchanger and

cause liquid condensation to be formed in the cavities of the plates, the

exchanger was positioned vertically with the exhaust air circulating

through it from top to bottom (see Figure 13). Provisions were made to

absorb and measure the condensates.

Initial tests were conducted under natural outside conditions, and

the air was forced (blown) inside at outside temperature. Tests for

frost formation required low outside temperatures. Therefore a refrigerating

unit had to be attached to the forced-air intake circuit. But, then when

these tests lasted several days, frost formed upon the refrigerating unit

itself. To solve this problem the forced-air intake was made to circulate

in a closed duct without any input of outside air (see Figure lk)

U.2 Magnitudes Measured

Flow rates: On each of the two circuits, flow rates were measured

with the help of a diaphragm -inserted within the rectilinear portions of

the ducts.

Pressure losses: Pressure valves were mounted in both circuits on

both sides of the plates. Continuous recording of pressure losses was

possible.

Temperatures: Thermocouples placed at various points within the

circuits were used to measure temperatures: they could be recorded on

multichannel potentiometers.

Air temperatures were measured particularly at the following points: —on the outside, —at the forced-air duct inlet

16
—at the exchanger inlet on the forced-air intake duct

,

—at the exchanger inlet on the exhaust duct

,

—downstream from the heating elements on the exhaust duct, —at the exchanger oulets on both circuits, immediately

where both streams issue from the plates. Since temperatures

were not uniform at these points, thermocouples were mounted

on frames forming on integral part of the exchanger, each

frame carried nine thermocouples (see Figure 15 )

,

—after mixing of the air streams , at the exchanger outlet

on both circuits.

k.3 Test Results

4 . 31 Pressure Loss

Measurements were made at various temperatures and for different rates

of flow. Figure l6 shows average results observed.

Although pressure loss did rapidly reach a high value for the higher

flow rates, it can be seen that for a prototype-exchanger flow rate of,

for example, 350 m-^/h, the pressure loss was some 120 Pascal, which is not

excessive.

Measurements were also taken of that pressure loss caused by partial

obstruction of the inlet or the outlet of the plates , and also by a dual

partial obstruction at both the inlet and outlet as depicted in Figure ITResults of the measurements are plotted on Figure l6.

The advantagesof such obturation are discussed in paragraph U.3^+.

k.32 Efficiency

The efficiency or effectiveness in question is solely that of the

heat transfer from one gaseous stream to another.

17
Where D. = mass rate of flow of exhausted air,

D

e = mass rate of flow of forced-air intake

tjj = pre-exchanger temperature of exhausted air,

t! = post-exchanger temperature of exhausted air,

t = pre-exchanger temperature of forced-air intake,

t' = post-exchanger temperature of forced-air intake,

we can define efficiency "by the equation:

For measurements, we must in principle observe the equality:

D, ft - t'J = De (t' e - t.)

In point of fact, that was not entirely the case. Actually, despite

the insulation of ducts on both sides of the exchanger, a heat loss

occurred in the distances covered between the plates and the point of

measurement (temperature after mixing of air streams). The equality shown

was not altogether verified. As a consequence the real efficiency is

underestimated if equation 1 is used. An estimate closer to the true

thermal efficiency can be given by the equation:

~D.(f.-f.) + DAU-tU) (2)

Efficiencies calculated with both equations, were D^ and D are

equal, have been plotted on Figure 18.

It appears that although r and r' did differ appreciably at low flow

rates, they were, on the other hand, practically equal at flow rates above

350 m

3 /h.

Maximum efficiency was attained at flow rates of the order of

1+00 m /h, where it amounted to 70 per cent.

18
The varying difference between r and r' as a function of flow rate

is explained by the fact that although stray (infiltration) heat losses

increase as flow rate rises, they increase relatively less than the

exchanger's own heat transfer.

The measured efficiencies did not vary appreciably with forced-air

intake temperatures.

Measurements were made in the case of different D n

- and D flow rates, i e

'

but these were not of considerable interest. In fact, when D^ is lower

than D , in other words when more air is forced (blown) in than is

exhausted, efficiency will be greater than that obtained when D is equal

to Dj_. It will even tend toward 1 when D^ is much lower than D

e

, but in

that case the exchanger might be of limited advantage. In the opposite

case, the efficiency is less than for two equal flow rates and tends

toward zero if the difference in flow rates is very pronounced.

H.33 Increased Efficiency when there Is Condensation of the Water-Vapor

in the Exhausted Air.

During the second series of tests , the exhausted air was humidified

by means of the apparatus described earlier.

Although the tests without humidification displayed a certain

constancy of the results vis-a-vis outside conditions, the same is not

true when—as may well be imagined—the following parameters are involved: —exhausted air's moisture content, —temperature of forced-air intake.

Only a limited number of tests were conducted because of the length

of time required to get the humidifying equipment operating normally.

These tests did reveal, however, that for exhausted-air relative humidity

19
below 50 per cent and for outside temperatures higher than -5°C, results

obtained did not differ appreciably from those observed in tests without

humidification.

For higher inside humidity conditions, an improvement was noted in

efficiency. Although limited, this improvement could be of three to four

per cent in the extreme case of a 100 per cent relative humidity.

It cannot be said, therefore, that in such an exchanger there was a

spectacular recovery of the "latent" heat corresponding to the water-vapor

content of the exhausted air.

On the other hand, it must be noted that the average relative humidity

encountered during the heating season in dwellings is not too high. And

even if there is a humidifier for the forced-air intake, it would be

economical to have that humidification accomplished at least possible

cost, and thus have it correspond to the minimum compatible with comfort

requirements. Moreover, it would be particularly illogical to employ in

one and the same building a heat-recovery system and a humidifying system,

in such a way as to generate waste

MOTE: It may be asked whether the definition of efficiency used

previously must still be accepted. The same equation may be used, but by

replacing temperatures with enthalpies. This is tantamount to including

within the amount of recoverable heat, that latent heat of vaporization

corresponding to the difference between the water-vapor content of the

exhausted air and the forced-air intake

Calculation of "enthalpic" efficiency evidently leads to a result

inferior to the one obtained with temperatures. This stems from the fact

that condensation of the water vapor contained in the exhausted air occurs

20
in only a limited part of the exchanger.

Let us consider an example that will he hetter understood by consulting

the psychrometric chart in Figure 19- Let us assume that at the exchanger

inlet the exhausted air is at 20°C and contains five grams of water vapor

per kilogram of air (point 1 on the diagram); its relative humidity is

thus 35 per cent and its enthalpy is 7-8 kcal/kg. Let us assume that the

forced-air intake is at 0°C and contains three grams of water vapor per

kilogram of air (point 2 on the diagram) ; its relative humidity is thus

80 per cent and its enthalpy is 1.8 kcal/kg.

Let us assume that the forced-air intake leaving the exchanger is

li+°C and the exhausted air is 6°C ; as the weight of the vapor in the

exhausted air has not changed—there can be no condensation for it, since

it is being warmed—its vapor weight is still three grams per kilogram of

air (point 2' on the diagram) and its enthalpy is 5.2 kcal/kg.

Therefore efficiencies are: —in temperatures:

t'.-t. 14-0 14 ?0%

U - te ~~ 20 - 20 —in enthalpies

:

r -

E '«- E« =5- 2 ~ 1 " 8 f- = 5d = 60 % E, -E. 7.8-1.8 6.0

/0

But this enthalpic efficiency is of no interest to us since the latent

heat cannot be utilized (to do so condensation would have to be accepted

in the dwelling). We shall therefore discuss only efficiency in

temperatures.

21
We should like to add that, by definition, in the case of equal flow

rates, efficiency in enthalpies is limited in each case to a number below

1 since there is no way of directly restoring water vapor from one medium

to the other and that the temperature of the forced-air intake after it

has passed through the exchanger cannot evidently be greater than the

temperature of the exhausted air. This statement does not apply in the

case of unequal flow rates. As an example, when the flow rate of the

forced-air intake is much higher than the flow rate of the exhausted air,

and if the relative humidity of the outside air is 100 per cent, it is

not theoretically impossible to reach an efficiency of 1 in "enthalpy"

recovery. But in practice is this a very interesting case?

k.3^ Supplementary Tests with Partial Obturation of the Exchanger Air

Inlets and Outlets

As we have shown in Figure 17, exchanger outlets and inlets can be

arranged to lengthen the average distance traveled by a fluid particle.

Furthermore, this arrangement can be such that the circulation of the two

streams is no longer purely a crossflow, but can at certain points be

assimilated to a counterflow circulation.

The value of the additional pressure losses resulting from this

arrangement were given in paragraph U.31. The observed gain in efficiency

is some three to five per cent, depending on the flow rates used.

The economic advantage of such an arrangement can be made evident

only by an "optimization" calculation.

k.k Study of Frost Formation

Appearance and Expansion of Frost

22
When the outside temperature is very low, an area appears in the

exchanger where the temperature of the plate surfaces on the extraction

side is much lower than 0°C . In these circumstances, and despite the

highly turbulent character of the flow, water vapor in the exhausted air

and condensed in suspension may deposit on the surface in the form of

frost. If no change is made in the equipment's operation, a block of

frost starts forming at the coldest point and then spreads out until the

exchanger becomes clogged (see Figure 20). It is obvious that the risk

involved in this instance is such that the colder the climatic region of

the exchanger location, the greater the necessity to take preventive

measures.

Factors Having an Effect on the Occurrence of Frost and Its Rapid

Expansion

A priori these factors are: —temperature of the forced (blown)-air intake (outside temperature), —temperature of the exhausted air, —humidity of the exhausted air

,

—flow rates.

We shall indicate the qualitative effect of each one of these factors ,

as observed in the course of our tests. —Outside temperature: occurrence of the frost-formation process is

linked to this temperature. Theoretically, there is danger of frost

forming whenever the temperature dips below 0°C . In actuality the

observed threshold is rather of the order of -3°C. It should be noted that

in a real case this limit must be lowered even more because of the

existence of various impurities in the exhausted air.

23
The speed with which the process spreads is then an increasing

function of the deviation between 0°C and the outside temperature. — Temperature of the exhausted air: its effect is counter to the

preceeding effect. If the temperature of the exhausted air is 2^°C

instead of 20°C, frost begins forming only at an outside temperature of

-5°C instead of -3°C

Humidity of exhausted air: the higher the humidity, the greater the

risk of frost. In humidity conditions like those normally encountered in

winter inside dwellings (30 to 60 per cent), this factor does, however,

seem less important than the others. — Flow Rates (the situation considered is one in which exhaust and

intake flow rates are close): the exhausted-air flow rate has no

perceptible effect on the appearance of frost. On the other hand, the

spread or expansion of the process is directly proportional to that flow

rate.

Effects of Frost

The formation of frost has a highly unfavorable effect on the

functioning of the exchanger and the ventilation system in general. It

manifests itself in: —a drop in exchanger effectiveness, —an increase in pressure loss throughout the exchanger. This

loss may be very substantial and completely alter the operation

of the ventilation system. Furthermore, we note that the loss

may occur during an off-period in the expansion of the frost

mass because during that off-period the "density of the

frost" remains "free" and becomes too great to allow

24
sufficient cooling of the exhausted-air stream. This situation

arises when in order to counter major pressure losses other than

those of the exchanger, the ventilators (fans) in a ventilation

system have a pressure (load) several times greater than the

exchanger pressure loss. But this ohviously does not

constitute a feasibile "self-regulating" procedure.

Figure 21 shows that evolution of the pressure loss through the

exchanger for different flow rates.

Procedures Designed to Avoid Formation of Frost

The patent statement which follows this report enumerates several

solutions. At this point we shall mention only three of them:

The first solution considered was to equip the exchanger with a

sliding panel or shutter that would deflect the exhausted-air flow away

from the exchanger's coldest point (see Figure 22). Tests show that

this solution was not entirely satisfactory. It was therefore rejected.

A second solution involved preheating the forced-air intake before

it passed into the exchanger. In this way the temperature of that air

would be raised above the threshold of initial frost formation. The

economic aspects of such a method will be examined later in this report.

Notice that this procedure does have the advantage of not creating any

discontinuity in forced-air intake temperatures during the winter, as

the third solution described below does.

The annual fuel consumption required for this preheating is very low.

In the Paris area, during an average winter and with the preheating

raising the air temperature to 0°C , this consumption represents only one

per cent of the total heating requirement.

25
A third solution consists in providing an exchanger by-pass so that

the cold forced-air intake does not flow through the exchanger when frost

is being formed. This device can be controlled by: —either a differential pressure controller ("pressostat" ) that

detects any abnormal rise in pressure loss throughout the

exchanger (see Figure 23), —or, in an open aperture, by a regulator "slaved" to the

outside temperature (a "pressotat" may be installed for

additional security)

This method has the disadvantage of introducing temperature

fluctuations in the forced-air intake which can cause condensation,

periodic expansions, etc. . . It may be noted that in a large-size

system there is still the possibility of introducing limited heating into

the by-pass. In this event, however, the advantage over the preceding

method is no longer evident.

U.5 Conclusion

The tests conducted on the prototype exchanger built at the CSTB

permitted us to draw certain conclusions or rules on the behavior of a

cross-flow exchanger of this type: —Its efficiency can be advantageous —of the order of TO per

cent —for pressure losses that are not excessive. —Disadvantages stemming from frost formation can be eliminated

by use of a suitable device.

Since the dimensions of the prototype exchanger were arbitrarily

chosen, we shall endeavor to determine those dimensions and flow rates

conducive to optimum exchanger utilization, by calculations based on the

26
data and results obtained experimentally.

5. EXTENSION OF EXPERIMENTAL RESULTS BY COMPUTER CALCULATION .

5.1 Calculation of Exchanger Efficiency

Let us take a crossflow exchanger and assign subscript i to the

exhausted air and subscript 2 to the forced-air intake. Temperatures t.

and tp at a point in the exchanger are a function of the X^ and X

? coordinates of that point, as defined in Figure 2k. When the flows are

laminar variations t, and t in space may be expressed by the differential

equations:

(1)

"A = * <* - «

U= kAh ~ h) -

where coefficients Is.-, and k_ are function: —of the exchanger characteristics: nature and shape of

plates, distance between them, —of the "flow-rate density" of the air stream being considered,

that is to say of the flow rate divided by the width of the

exchanger and the number of plates

.

—of the temperature of the air current.

Knowing functions k-, and k

? and the temperatures at the inlet to the

exchanger, tj_ and t , we can gradually calculate the temperature field in

each flow, and therefore temperatures at the exchanger outlet, and

therefore the efficiency. This calculation can be done quite easily with a

computer. A program has been written for this purpose at the CSTB.

27
In the practical example we have under study flows are not laminar.

But nevertheless, the above theoretical representation can be used with

sufficient approximation. Moreover, as temperatures remain within a

relatively limited interval, it may be assumed that k and k are constant

in space. And lastly, if the flow rates of both fluids are equal, and if

the exchanger is square, by symmetry we have equality between k-, and k_.

We shall assume that situation for this computation.

Efficiency is expressed:

lj o h (Xi. L) efri - l J fl h (xlt 0) dxj,

1 fL I f

h (0. xj dx2 ~ijo h (*i. 0) dXl

where L = width of the exchanger.

t-, (0, X ) and t (X, , 0) are constants equal to t and t

respectively, so that the above expression may be written simply as:

1 fL

, L) dxx - te

For a given value of k, r depends only on L. Figure 25 shows the

variation of r as a function of L for k = 8 m~ . This variation was

determined by computer calculation with the program mentioned above.

It follows from the form of the differential equations (l) that if

we know an r

Q function (L, kQ

) of r as a function of L, we know every r

function ( L, k) by the transformation:

r(l.k) =r (L r k ).

By applying the method of computation just outlined, it is possible

to determine the efficiency of an exchanger of any dimensions, when knowing

the efficiency, determined experimentally, of an exchanger fabricated with

28
the same plates. We made this determination for our type of exchanger.

The result presented in Figure 28, is depicted by two curves. The first

gives the variation of efficiency r as a function of the product A X L/L »

with LQ being the length of the prototype, L the length of the exchanger

being considered, and A the coefficient function of the ratio of the

"flow-rate densities," d/dQ

, of the two exchangers (let us recall that the

density of the flow rate is equal to the total flow rate D divided by the

width L of the exchanger and by the number n of its plates). The second

curve gives the variation of A with d/d .

5.2 Calculation of Pressure Loss

Pressure losses at the inlet and outlet of the exchanger are negligible

when viewed against the losses that occur along the distance covered

between the plates. Pressure loss through the exchanger is consequently: —proportional to the width L of the plates

;

—function of the flow-rate density d; —measurements showed that this function is essentially of

2 form d .

Therefore, between pressure losses hQ and h of two exchangers of L

and L sizes and flow-rate densities d and d, there is the relation:

This provides a means of calculating the pressure loss through an

exchanger of the same nature as the prototype exchanger, but of different

size and flow-rate density.

In the next chapter we shall show how the equations just established

for efficiency and pressure loss will enable us to calculate the optimum

29
dimensions of the exchanger under given conditions.

6. ECONOMIC AND INDUSTRIAL ASPECTS

We shall first compute the savings in heating achieved by use of an

exchanger, and also calculate that exchanger cost which must not be

exceeded if it is to be a cost-effective proposition.

We shall then show how to determine, for given conditions, the

optimum dimensions of an exchanger whose prototype has previously been

laboratory tested.

6.1 Computation of Savings in Heating Obtained with an Exchanger, and the

Latter 's Cost-Effectiveness Threshold .

Although the presence of an exchanger does reduce heating costs, it

does, on the other hand, entail an increased pressure loss which in turn

increases ventilation costs. The reduction in heating costs and the

increase in ventilation costs must be computed for both initial investment

costs and operating costs.

6.11 Savings in Heating

We shall compute the savings for one cubic meter per hour of incoming

air passing through the exchanger.

The investment savings is equal to c(t. - t )i, where c = specific

heat of the air, or 0.3 kcal/°C ; t- = basic inside temperature, in

dwellings generally taken as equal to 20°C; t = basic outside temperature

whole value depends on the building site and location; i = marginal price

of the installed kcal/h.

Operating savings are equal to c 2k D . f where c = specific heat of

J

30
the air, already mentioned above; D. = value of the degree-days of heating

J at the building location; f = maximum price of the consumed or expended

kcal.

Values of i and f are function of numerous factors: type of heating

(individual, collective for a multi-story (apartment) building, collective

for a group or block of houses and/or buildings), fuel, boiler-room or

furnace-room facilities, flexibility and regulation. . . Our computations

will be for average values which we will take as equal to: —O.k francs per installed kcal/h (i) —0.00003 francs per consumed kcal (f)

And lastly, we shall compute for three pairs of values of t and D^

corresponding to about the average of these values for each of the

climatic zones, A, B, C, defined in the CSTB Technical Report:

Zone A: t

Q = -15°

Zone B: t = -10°

e Zone C: t

g = - 5°

The savings obtained under these conditions are as follows:

Investment Savings Operating Savings

Zone A 1+.20 francs 0.60 francs per year

Zone B 3.60 francs 0-50 francs per year

Zone C 3.00 francs O.Uo francs per year

6.12 Cost of Increased Ventilator (fan) Capacity

The investment and operating expenditures entailed in augmenting the

capacity of the ventilator fans depends, as in the case of heating

savings , on various factors , and first and foremost on the exchanger

31

D

3 = 2800

D

j = 2300

D

o = 1800
pressure loss (the expenditures may be estimated as almost directlyproportional to the loss) . But fortunately these expenditures remain

relatively low when compared to the savings in heating. Therefore, in the

computations made in this paragraph we shall take only one value, namely

the one for a pressure loss of the order of 200 Pascal. This is the loss

that exchangers designed after our prototype should have. The figures

retained are as follows (per m /h)

—in investment expenditures: 0.25 francs —in operating expenditures: 0.10 francs per year

6.13 Summary Analysis

Table 1 is a compilation of: —the values given above, i.e., for savings in heating and

for expenses incurred in augmenting ventilator-fan capacities,

per nrVh; —the difference between these savings and expenditures, still

per m /h; —operating savings expressed as investment savings for an

interest rate and amortization period of 10 years: Doing

this involved multiplying by the factor:

1 1 1 + + n , mMO = 6- 1 5

(1 + 0,1 ) ^ (1 + o,i ) 2 ^ ^ (1 + 01 ) 10 — overall savings expressed as investment savings, still

per m /h; —values obtained when the preceding figure is multiplied

either by 180 or 3600 (the former gives the approximate

savings per dwelling, the latter per 20 dwellings /apartment) .

32
To be sure these are approximate figures, but, nevertheless, they do

fix the order of magnitude of that exchanger cost ceiling which must not

be exceeded if use of the exchanger is to be profitable. The following

expenses must be included in that exchanger cost: initial investment,

maintenance and service, replacement of plates if the "expendable"-plate

solution is selected.

Our estimates seem to show that there should be no difficulty in

staying below the cost-effectiveness or profitability threshold we have

just computed. Table I Zone A Zone B Zone G 5 Savings in heating per M /h

in investment costs

in operatings costs

4.20 francs

0.60 francs

per year

3.60

0.50

3.00

0.40

Expenses due to increased capacity

of ventilator fans per ¥r/h

in investment costs 0.25 francs

in operating costs 0.10 francs

per year

0.25

0.10

O.25

0.10

Differences between savings

and expenditures per Jr/h

in investments

in operations

3.95 francs

0.50 francs

per year

3.35

0.40

2.75

0.30

Operating savings expressed

as investment savings 3.05 francs 2.45 1.85

Overall savings expressed

as investment savings

per M

3/h ~ for 180 M

J/h (one

dwelling)

for 3600 M-yh (20

dwellings or

apartments)

7.00 francs

1260 francs

25,200 francs

5.80

1045

20,900

4.60

830

16,600

33
Our computation did not take the forced-air intake duct into account.

We assume that such a system is required for example, because the dwelling

in question is in a noisy area (windows cannot be opened). However, if

this duct cost is taken into account, then we find that it is difficult

if not impossible to make the exchanger a profitable proposition. Use of

the exchanger therefore implies that a mechanical forced-air intake is

warranted for some other reasons, at least in part. In such a situation

the employment of the exchanger may be the additional determining factor.

6.2 Method of Computing Optimum Exchanger Dimensions

In general, there are two types of exchangers worthy of consideration:

1. Exchangers designed to handle high flow rates (multipleoccupancy buildings). In this case, the pile of plates may consist of

an assembly of a certain number of standardized basic modules.

2. Exchangers designed for an individual dwelling. In this case the

most attractive choice appears to be a single compact unit containing the

ventilation fans, the exchanger with its anti-frost system, and the filters

if these are provided.

For both types , the manufacturer would find it advantageous—if only

to facilitate the replacement of plates—to offer a very limited line of

units or modules. Thus the manufacturer should determine the optimum

dimensions of these units or modules. Then the user will select the optimum

unit or the optimum number of modules to fit his particular situation.

We shall show how such optimization computation may be made.

We shall consider the problem in its general aspects.

34
Given: flow rate D to be handled, and interest rate x and an

amortization period of N years. It is assumed that the exchanger selected

has square plates and a geometry identical to that of a prototype whose

performance characteristics have been determined by laboratory tests.

The problem is to determine the length L of the exchanger and number n of

plates conducive to the least overall cost (investment expense and

operating expense).

We propose to examine successively each of the elements entering

into these expenses so as to see which of them are affected by L and n

and, if so, how these characteristics do affect them.

6.21 Savings in Heating

Savings in heating, with respect to both investment and operating

costs, is proportional to the efficiency of the exchanger. As we saw in

paragraph 5.1 that efficiency depends upon length L and flow-density d.

Since for a given flow rate d is inversely proportional to L and n, in

the final analysis efficiency depends only on L and n. In paragraph 5-1

we described how it was possible to calculate r as a function of L and d,

thus of x and n.

And in paragraph 6.11 we showed how, knowing the efficiency of the

exchanger, the savings in heating costs could be computed.

6.22 Cost Due to Increased Ventilator-Fan Capacity .

We may assume, with sufficient approximation for this computation,

that the cost—investment and operating—of the increased capacity of the

ventilator fans is proportional to pressure loss.

In paragraph 5-2 we saw that this pressure loss is itself proportional

to the length L of the exchanger and to the square of density of flow d.

35
As, for a given rate of flow, d is inversely proportional to n and L,

o pressure loss is finally inversely proportional to Ln .

And in paragraph 6.12 we gave the order of magnitude of the

investment and operating costs for an exchanger pressure loss of 200

Pascal.

6.23 Exchanger Cost

We may consider two parts in the exchanger:

1. The plates which in our preferred solution are to be replaced

annually. Their cost should be counted as an operating cost. The cost

of a plate is divided approximately into a fixed part and a part

proportional to the square of the length L of the exchanger. Therefore,

for n plates we have a cost equal to n(a + bL ) , where a and b are

constants

2. The rest of the exchanger (case, defrosting system) and its

connection to ventilation circuits. It may be assumed as a first

approximation that this cost is independent of exchanger size (the cost

approximately breaks down into a fixed part and a part proportional to

flow rate)

A few operating costs should be added to the above (maintenance and

service, operation of the defrosting equipment is so equipped). These

costs are practically independent of the exchanger dimensional

characteristics L and n.

6.2k Summary

We have not, for the moment, worked with a manufacturer on this matter.

We therefore do not have sufficiently precise data on the price of the

exchanger or plates to be able to present an optimization computation.

36
But the particular purpose of this paragraph was to show how such a

computation can he done.

7. CONCLUSION

The solution studied at the CSTB for a system designed to recover

heat from ventilating air, appears satisfactory from the sanitary and

hygienic aspect and also from the viewpoint of economy and savings.

APPENDIX — COMPUTATION RELEVANT TO HUMIDIFICATION OF THE AIR IN A DWELLING

By following the psychrometric chart in Figure 19 > we shall consider

a day during which the weight of water vapor in the air is two grams per

kilogram of air. Taken at 21°C this gives a relative humidity of 13

per cent. But inside a room there is a certain reduction of the water

vapor which we may estimate to he approximately 30 grams per hour,

providing the dwelling is properly ventilated. If the air change in this

•3 "3 room is 30 m°/h, this means that the air in that room contains per m ,

i.e., as a first approximation per kg, one gram more than the outside air.

Therefore , there will he 3 grams per kilogram of air which corresponds to

a relative humidity of 20 per cent.

To ohtain 35 per cent, that is 5-3 grams per kilogram of air, requires

the evaporation of 2.3 grams per m^ per hour, or approximately 1.7 liters

for 30 m-* and 2h hours .

37
The figures given in paragraph 2 of Chapter 1 (introduction), result

from computation and statistics compiled from meteorological data

obtained at Paris-le-Bourget. The data show that on an average (average

winter i960 to I96U) there are: —8 days during which the weight of the water vapor is

between one and two grams

—20 days when it is between two and three grams , —37 days when it is between two and four grams, —hk days when it is between four and five grams.

The relative humidity in the room under consideration is equal to 13,

20, 26, 32, and 39 per cent for respectively 1, 2, 3, 4, and 5 grams of

water per kilogram of outside air. Extrapolation of these figures shows

that the relative humidity in this room will be

—below 20 per cent during 8 days

—between 20 and 25 per cent during 16 days, —between 25 and 30 per cent during 28 days, —between 30 and 35 per cent during 30 days

Conversion Factors

1 ft = 0.305 m = 30. U8 cm

1 m =39.*+ in = 3.28 ft = 1. 1 yd

1 cm = 0.39 in = 0.033 ft =0.011 yd

1 m

2 = 10.8 ft 2 = 1.20 yd2 1 ft 2 = 0.09 m

2 =0.11 yd2 1 yd2 = D.Qh m

2 =9.0 ft 2 1 m

3 = 35-3 ft 3 = 1.31 yd3 1 ydJ =27.0 ft J = 0.76 m

3

Pa = Pascal, Newton/m2 1 Pa = 0.00^01^7^ in water gauge pressure

°C = (F -32)/l.8

1 kcal/m'Vc = 0.20U8 Btu/ft2h°F

1 m3 /h = 35.3 ft 3 /h

38
"i

<D +-

(0E

U)

a. i O

in -0

roI H101

0)1

x

—

ixi ro

:<>v ...

+-

roc

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in l_o

M- (/) n J_>10 F 1 ID +-

!_

m

CD 0)

ro j_ TD ro * !_ o CD — >

u_

O

_j

0)1_

39
Figure 2. Sketch of exchanger with

therma I wheel

Batteries

Exhausted afr

Forced-air intake

Figure 3. Exchanger using an Intermediary

heat-carrying fluid.

40
s_O

+-

<o

!_o

-Q

(U

-a

if)

to

s_

toQ- <

S_

=1CD

41
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Figure 5

^tmsmmma&M^^fa $M

Figure 6

Figure 7

42
Exploded view of

a corner assembly

Parts to be sealed at each level

Detail view of stacking of

plates at the level of the corners

Adhesive-bonded edges

Figure 8. Assembling of plates

Pile of plates Polystyrene

-Wooden corner 40 X 40 X 202

Figure 9. Exchanger unit

43
F i g u re 10

Figure I I

44
CD

(0 -CUX

CD

CDCOE

(D

CD E

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01 +-o

45
46
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47
tf) uOu

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4-

ro!_Q

Qujuu QfrZ uuiL6L

48
(0u

(0 Q.

50 1D0 150 200 250 300 350 400 Flow rate i n m3/h

Figure 16. Exchanger pressure loss as function of flow rate

a) obturation only at

outlet

_ forced-air Intake (cold)

-- exhausted air (warm)

b) obturation at both outlet and inlet

Figure 17. Diagram of partial obturation of exchanger

49
>-u

T» BO 200 250 300 350 «0 450 500 550 Flow rate tn mVh

Figure 18. Variation of exchanger efficiency as function

of flow rate, when exhaust flow rate and forced-air

intake flow rate are equal.

50
feft 10

5 10

Temperature Jn °C

Relative humidity

Figure I 9.

51
Forced-air Intake

Exhausted air

Coldest point where

frost begins to form

Figure 20. Point at which frost forms

52
CD

Pressure loss In Pascal 1

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L-

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LD 1 CM fO

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Temperature In °C o ro O s_ o ro

LO

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53
Pressure loss In Pascal

M O-J '

\

\

\ -- . J - — P H —

\ r

\

\1 \

^

L_,

I —— ——

\

'

ifi\ O)

i

3\

1 <0\

-M

***! t>

(O

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EOo

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CL O

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Temperature fn °C

54
Forced-afr intake

Movab le shutter

Figure 22. Diagram of the anti-frost solution considered.

55
pressure tube

Figure 23. Automatic by-pass defrosting system

The principle of this s limits are set on the p pressure loss increases

pressostat. The latter

air intake into the byby the forced-air but i frosts the exchanger, t limit set on the presso

shuts off the by-pass a normal flow through the

ystem is as follows: The two desired pressure loss

ressostat (Pressurestat) . When frost forms, the

until it reaches the upper threshold set on the

actuates the 3-way valve which diverts the forcedpass circuit; thus, the exchanger is no longer cooled

s warmed by the exhausted air. This warm air dehus lowering pressure loss down to the lower threshold

stat. The latter reactuates the 3-way valve which

nd thus allows the forced-air intake to revert to its

exchanger.

56
J

L

X

i , J i i i

I i

, i

(te )

, i L «-

"tf

( ti) L

xi

Figure 24,

100

90

80

feSi. 70 c— GO >-O

c

50

40 U

>t- 30

LU

20

10 __l 1 1 1 1 1 1 1 1 1 to

10 20 30 40 50 60 70 60 90 100 Length L in cm

Figure 25. Variation of exchanger efficiency as a function of exchanger length L, with construction

of plates and rates of flow such that kj = k2

= 8.m"

and when number of plates is equal to 1 00.

57
Effici ei c y

i

/

«

• fa lue of 'a X l/l'

c

5 2

J€ Df A

Value of a/a

o

Figure 26. Variation of exchanger efficiency

as function of its characteristics.

58
FORM NBS-114A (1-711

1. PUBLICATION OR REPORT NO. NBS TN-710-52. Gov't Accession No.

U.S. DEPT. OF COMM. BIBLIOGRAPHIC DATA

SHEETS

3. Recipient's Accession No.

4. TITLE AND SUBTITLE BUILDING RESEARCH TRANSLATION

USE OF AN AIR-TO-AIR HEAT EXCHANGER TO RECOVER HEAT FROM AIR EXHAUSTED BY MECHANICAL VENTILATION

5. Publication Date July 1972

6. Performing Organization Code

7. AUTHOR(S)

P. Garrivier

8. Performing Organization

10. Project/Task/Work Unit No. U21 31.1k

9. PERFORMING ORGANIZATION NAME AND ADDRESS NATIONAL BUREAU OF STANDARDS DEPARTMENT OF COMMERCE

WASHINGTON, D.C. 20234

11. Contract/Grant No.

12, Sponsoring Organization Name and Address Same as No. 9-

13. Type of Report & Period Covered Final

14. Sponsoring Agency Code

15. SUPPLEMENTARY NOTES

16. ABSTRACT (A 200-word or less factual summary of most significant information. If document includes a significant bibliography or literature survey, mention it here.)

When a building is equipped with a mechanical system which both exhausts and supplies

outdoor air, it may be profitable to install a heat exchanger between the two circuits. The CSTB has perfected such an exchanger, which meets the two principal requirements

of the problem: good efficiencies and elimination of the polluted air intake. A method of calculating the optimal dimensions of the exchanger has also been devised.

17. KEY WORDS (Alphabetical order, separated by semicolons) Exchanger dimensions; frost formation; heat exchanger; heating costs; mechanical

systems; pollution; pressure loss; ventilation.

18. AVAILABILITY STATEMENT

r~Xi UNLIMITED.

I I FOR OFFICIAL DISTRIBUTION. DO NOT RELEASE

TO NTIS.

19. SECURITY CLASS

(THIS REPORT)

UNCLASSIFIED

20. SECURITY CLASS

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21. NO. OF PAGES

62

22. Price

$.60

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