Residential space heating with the heat pump

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Residential space heating with the heat pump

20.38

RESIDENTIAL

SPACE HEATING WITH THE HEAT PlFMP by

R.L.D. Cane

In Canada

today, o i l and n a t u r a l gas

a r e

by

far

t h e most commonly used fuels for residential space heating. In 1976, oil and gas

h e a t i n g systems were installed in 85 per c e n t of the e x i s t i n g housfng stock.' Electric heating systems of one form or another accounted for

o n l y 13 per cent: about the same proportion as a f t e r their introduc-

tion in the l a t e 1950's. In s p i t e of a relatively low i n i t i a l cost when compared with gas and oil h e a t i n g equipment, electric h e a t i n g

was and still is expensive

from

an operating-cost paint

of

view. In

many regions of the country t o d a y , however, the differential between convenrional fossil-fuel and electricity p r i c e s i s narrowing and electricity i s being used to a greater extent t h a n in the past. One

d e v i c e , which utilizes electrical input to produce heat and promises

to make electrical h e a t i n g more economical from an operating-cost and

prime energy-use standpoint, is the heat pump.

The heat pump employs the same basic principle as the common

household refrigerator: heat is extracted from a space at l o w temper- ature and discharged to another space a t a h i g h e r temperature. In

fact, any process that results in rransfer o f h e a t from a law temper-

a t u r e t o a h i g h e r one i s referred t o as "heat pumping." By this

definition an air conditioner and a r e f r i g e r a t o r are b o t h heat pumps because h e a t i s removed from a conditioned space and discharged to t h e h i g h temperature s u r r o u n d i n g .

The h e a t pump employs the vapor compression refrigeration cycle

in order to t r a n s f e r heat from a body a t a Xow temperature to one at a higher temperature (Figure I ] . The b a s i c components employed in

the cycle are two heat exchangers, compressor, expansion valve and interconnecting p i p i n g . A fluo~ocarbon refrigerant such as Freon 22

is usually employed as t h e working f l u i d in the cycle. The temperature

at which the l i q u i d refrigerant vaporizes can b e controlled by reducing

the

pressure in the evaporator. Similarly, by increasing xhe pressure

of t h e vapor in the compressor, the temperature is raised above the

temperature of t h e sink and the vapor condenses, liberating b o t h

l a t e n t heat and heat of compression. The h i g h temperature, h i g h

pressure liquid then passes t h r o u g h an expansion device, which reduces

the pressure o f the liquid to the pressure

in

the evaporator s e c t i o n and the cycle is repeated.

The

unique

advantage of t h i s cycle i s that t h e energy obtained

from t h e condenser as heat

can be

appreciably greater than t h a t

associated with the electrical power r e q u i r e d to drive the compressor.

The measure of performance of such a device; called the coefficient o f performance (COP], i s defined as t h e ratio of t h e d e s i r e d e f f e c t t o t h e heat equivalent of compressor i n p u t . It i s always greater

than I . Men operating on the cooling c y c l e , t h e d e s i r e d e f f e c t i s

t h e h e a t t a k e n in at the evaporator; when operating on t h e heating

c y c l e , t h e desired e f f e c t

is

the heat liberated by _{t h e condenser. By}

reversing t h e d i r e c t i o n of the refrigerant flow through the heat

exchangers, t h e heat pump can provide e i t h e r heating

o r

cooling as

required (Figure 23,

AIR-TO-AIR HEAT PUMP

The most widely employed h e a t pump uses the outdoor air as a

h e a t source or sink [Figure 23. In the h e a t i n g mode, outside a i r i s

dram through t h e evaporator h e a t exchanger, gives up some heat to

t h e vaporizing r e f r i g e r a n t , and is discharged from the outdoor s e c t i o n . The compressor increases the pressure and temperature af t h e refrigerant

vapor allowing it to condense in t h e indoor h e a t exchanger where i t s

heat i s transferred to the air from the room. The heated air is t h e n

returned to the room.

The air-to-air heat pump was originally introduced in the s o u t h e r n

United States, where cooling and dehumidification were of prime impor-

tance and h e a t i n g seasons were short and mild. The units were sized

to cape with the cooling l o a d and in most applications such u n i t s

a l s o provided s u f f i c i e n t h e a t i n g capacity. However, heat pumps marketed

in more n o ~ t h e r n regions had t o be equipped with supplementary

electrical resistance heaters. This is still t h e practice today. PERFORMANCE

OF

AN AIR-TO-AIR HEAT PUMP

Unlike conventional h e a t i n g systems u s i n g natural gas or oil, the

h e a t i n g capacity of an air-to-air heat pump depends on the outdoor

temperature [Figure 3 ) . Both the capacity and COP of the heat pump

a r e reduced as t h e outdoor temperature drops. The outdoor temperature

at which the heat pump can j u s t satisfy t h e heating requirement of

The space is referred to as the "balance" point. The balance p o i n t is a function of the building heat loss and t h e capacity of the heat pump.

Below the balance p o i n t t h e hear pump has i n s u f f i c i e n t capacity t o meet the demand, and t h e d e f i c i t is made up by supplementary

heaters. The supplementary heaters are staged to come on in such a

manner that o n l y t h a t portion actually r e q u i r e d to make up the d i f f e r -

ence between the heat pump capacity and t h e space h e a t i n g requirement i s a c t i v a t e d . The COP, including the electrical input to the heaters,

heaters are brought

on-line.

Above t h e balance point t h e opposite is true; the u n i t h a s surplus capacity and

is

madc t o match t h e building heat l o s s by c y c l i n g on and o f f .

It i s i n t e r e s t i n g t o examine the performance of a

heat

pump on a

seasonal b a s i s . With t h e following i n f o r m a t i o n , one may estimate the contribution o f the h e a t pump t o t h e total h e a t i n g requirement:

(i] the manufacturera s performance data (i

.

e . , capacity, input power, and COP versus outdoor temperature),

[ii] t h e building heat loss characteristic,

(iii] the frequency o f occurrence of outdoor temperature requiring

h e a t i n g

f o r

the area of interest [<18"~; < 6 5 " ~ ] .

As an example, consider a h e a t pump with a nominal cooling

capacity of 7 kW (2 tons) and a

n e t

heat loss of 8 . 6 k W at -26°C I - 1 S 0 F ) as shown

i n

F i g u r e 4 , f o r a home

i n

t h e Ottawa area. The balance p o i n t is approximately - 1 ' ~ ( 3 0 ' ~ ) . The cross-hatched region

in t h e f i g u r e

is t h e heating requirement made up by t h e supplementary heaters.

Figure 5 shows t h e frequency distribution o f hourly dry bulb temperature f a r intervals of 3 " ~ ( 5 " ~ ) . This chart is based on an average far a 10 year p e ~ i o d between 1953-1966 f o r Ottawa International A i r p o r t .

The next s t e p is to determine average heat

loss,

heat pump

input and output, and t h e supplementary heat requirement (Table 1)

f o r each of t h e temperature intervals. F o r this example, t h e seasonal

performance f a c t o r (SPF) calculated is 1 . 6 , which represents a 37 p e r cent k i l o w a t t - h o u r saving over s t r a i g h t e l e c t r i c a l resistance

heating. (The actual energy saving would b e somewhat less s i n c e the manufacturer" r a t i n g s a r e obtained a t steady-state full-load conditions, while in practice t h e unit cycles on and off above t h e balance p o i n t . ]

F o r illustrative purposes, t h e heating energy inputs to t h e

residence a t various outdoor temperatures were p l o t t e d (Figure 6 ) . The total area (A+BcC) approximates t h e annual h e a t i n g energy require-

ment o f t h e house used in the example. Area A

*

B is a f r a c t i o n of t h e annual h e a t i n g energy requirement supplied by t h e heat pump. Area

0

is

t h e energy input t o the h e a t pump required t o r u n the compressor

and f a n s . Area C is t h e supplementary h e a t i n g energy supplied to the house and area R i s a f r a c t i o n of the total h e a t i n g energy obtained from t h e outdoor a i r (-37 per c e n t ) .

POSSIBLE PERFORMANCE

IMPROVEMENTS

To

the right of the balance p o i n t the heat pump h a s excess capac- ity relative to space h e a t i n g demands. If t h e h e a t pump capacity

could be matched to the net building heat l o s s above t h e balance p o i n t ,

a l s o result in increased reliability because the effects of cycling

on and off are eliminated. In theory, this could be achieved by

u s i n g an i n f i n i t e l y variable speed compressor. In practice, a two-

speed compressor would improve performance significantly. In addition to t h e improvements possible through capacity modulation, there exist a number of areas where modifications will

improve h e a t pump performance in cold climate regions.

Depending on t h e relative humidity, frosting o r icing o f t h e surface reduces the

a i r

flow through the face of t h e c o i l when t h e outdoor a i r - t o - r e f r i g e r a n t coil is at or

below 0°C ( 3 2 ' ~ ) . The result is a gradual reduction in capacity t o the point where the heat pump must be r e v e r s e d

(to cooling mode) t o d e f r o s t t h e c o i l . The frequency o f defrost required is g o v e ~ n e d to a large e x t e n t by the f i n

spacing om the c o i l . The greater t h e number

of

fins per

inch, the more frequently the machine must be d e f r o s t e d .

By increasing the primary surface area [tube area) and

r e d u c i n g the secondary surface area (fin area), a reduction in d e f r o s t c y c l e s should be p o s s i b l e .

In present split-system heat pumps (one part i n s i d e and

the o t h e r outside), the outdoor section contains the

compressor, outdoor coil, r e v e r s i n g valve, outdoor fan motor

and fan. Liquid and vapor refrigerant piping interconnects

indoor and outdoor sections. Heat losses from t h e vapor line, reversing valve and l i q u i d line could be reduced by moving

h i g h temperature components indoors. In some designs t h e

condensed refrigerant is cooled below the condensing temper-

ature by l e a v i n g the liquid

line

uninsulated.

This

i s

obviously a cempromise

in

favour of t h e cooling mode because

it has always been central air-conditioning practice t o

locate the compressor, condenser, fan and motor outdoors to

readily dissipate heat to the environment. This heat could be put to b e t t e r use to improve the h e a t i n g cycle.

ECONOMICS OF

THE

AIR-TO-AIR HEAT PUMP

It has been generally accepted that a heat pump could n o t be

j u s r i f i e d economically unless it could b e used f o r summer cooling. In Canada, the need for central air conditioning may not be a major

factor, therefore t h e decision whether t o use a heat pump s h o u l d be

based on h e a t i n g energy savings alone.

Thc installed c o s t o f an a i r - t o - a i r heat pump, suitable f o r

residential space h e a t i n g v a r i e s from $2000 t o $3500 depending on

capacity ( 5 , 3 t010.5 kW) (1 1 J 2 t o 3 ton). T h i s assumes t h a t both t h e duct work and electrical service will n o t have to be upgraded in the case o f a retrofit application. Consider the example where a central

c 1 c c t r i c f u r r ~ n c e and h e a t pump ;i rc h c i n g comporcd for t h c rcs i clrnc r?

mentioned previously. The d ~ i c t work and electrical scrvicc wo1iI3 l ~ c

t h e same in either casc. Thc comparative costs of t h e two systems

are shown in Table 2. I n t h i s i n s t a n c e , i s the h c a t pump a b e t t e r

investment t h a n t h e e l e c t r i c furnace? I f t h e u s e f u l service l i f e of

the heat pump is assumed to be P5 yeass, a constant annual s a v i n g of

$166 would only j u s t i f y an investment of $1262, assuming an interest

raze o f 10 per cent, whereas the e x t r a capital investment r e q u i r e d for a heat pump i s $1500. The u n i t c o s t o f e l e c t r i c i t y , however, has

been i n c r e a s i n g steadily in t h e past f e w years. This is expected to continue and should be taken into account in t h e analysis.

Assuming that the u n i t c o s t o f electricity escalates at a constant rate, the maximum investment t h a t can be justified is

determined by the equation :

1

Maximum lnverinenf durtiiicd = 5 x

-

_{l + i}

[s]

where S = first year saving

(J)

i = i n t e r e s t rate

l + e

a = - [e = electricity unit cost escalation rate)

l+i

n = useful service l i f e (years).

For i = 10 per cenl, e = 8 per cent, S = $166 and n = 15 years, the maximum investment j u s t i f i e d = 12.03

x

$166 = $3997. Thus the heat

pump would be a better investment than the electrical fnrnace i f t h e

e x t r a cost was l e s s than $1997. What effect would changes in t h e

v a r i o u s assumptions have on the analysis? Consider the following

cases :

(1) an SPF o f 1.4

rather

than 1.6 with other assumptions

t h e same. This would reduce t h e first year savings

to $1 17 (-30 per cent) ;

Maximum Investment Justified = $1407.

(23 an electricity u n i t energy c o s t escalation r a t e of 1 2

per cent and reduction in t h e SPF as in

(11,

with

o t h e r assumptions t h e same;

Maximum Investment J u s t i f i e d = $1815.

(3) as

i n

(13, b u t in addition a maintenance c o s t difference

of $50 per y e a r ;

Maximum Investment Justified = $1165.

A r e d u c t i o n in the SPF, from that calculated, would be likely, f o r the reasons o u t l i n e d previously, b . e . , because of on-off cycling

above balance p o i n t and defrost energy requirement, Case (31 raises

problem. Recent i n d i c a t i o n s are that manufacturers have solved a

number of problems that plagued early air-to-air heat pumps. The

average annual service c o s t reported

in

a recent study was approx-

i m a t e l y $SO per year. 2

Comparing case (1) and (2), one can see the importance of knowing

energy u n i t c o s t trends.

In

the case of e l e c t r i c i t y , t h e unit c o s t

escalation rate will, in all likelihood, exceed

the i n f l a t i o n r a t e on

o t h e r goads and services. Factors such as the utility generation mix,

i.e., the percentage f o s s i l f u e l , nuclear o r hydro generation, i s a

major factor detemining t h e wholesale unit c o s t increases and sub-

sequent r a t e increases t o the consumes.

ALTERNATIVES

TO

SUPPLEMENTARY

ELECTRIC RESISTANCE HEATERS

The SPF calculation of t h e air-to-air heat pump showed that t h e

supplementary h e a t e r s provided only 2 3 per cent of the total heating

requirement. This reduced the heat pump efficiency by 23 per cent

( i . e . , from a COP of 2.11 to a

SPF of 1.61).

More

significant is the

impact on t h e electric utility peak demand. The heat pump has t h e

same peak power demand as e l e c t r i c resistance heating, although it uses 37 per c e n t less kilowatt hours than the resistance h e a t e r s over

t h e course o f t h e heating s e a s o n . The utility must still provide

g e n e r a t i n g capacity based on peak demands.

In o t h e r wards,

heat

pumps have a poorer load factor f o r t h e utility than electrical

resistance h e a t i n g and thus would contribute to higher rates for electricity.

By eliminating t h e supplementary resistance heaters, t h e heat

pump would present t h e utility w i t h a desirable type af load because

input power

for the

compression c y c l e drops o f f a t reduced outdoor

temperature (Table 1). There are alternative ways of providing the

supplementary h e a t i n g . Today a number of manufacturers offer packages which may b e installed on a new

or

e x i s t i n g o i l or gas furnace. When

located in t h e p o s i t i o n

of

a normal cooling c o i l , t h e heat pump can

provide a l l the h e a t i n g above the balance point, while the furnace

provides the h e a t i n g below it.

One drawback to the heat pump furnace combination is that the

heat pump and furnace cannot be operated simultaneously (i.e., t h e

heaz pump indoor c o i l is located downstream of t h e furnace discharge).

The reason f o r this i s t h a t t h e temperature of the return air (air

e n t e r i n g t h e heat pump indoor coil) must be limited, otherwise

compressor power i n p u t increases resulting

i n

a h i g h r e f r i g e r a n t

d i s c h a r g e temperature.

This would

r e s u l t in a reduction i n performance

[COP) and h e a t i n g capacity, and would impose high stress conditions on

t h e compressor. Other methods allow t h e heat pump to run below t h e

balancc p o i n t . By employing a two-stage t h e r m ~ s t a t , ~ the furnace cycles on at the balance p o i n t temperature [sensed by a t h e r m i s t o r

located

in t h e

air stream entering the outdoor c o i l ) . Men the furnace

supply a i r

reaches

a predetermined temperature the heat pump cyclcs

o f f . A f t e r t h e furnace h a s

satisfied

t h e second stage and cycled o f f ,

t h e indoor blower

continues

to run until supply a i r drops to a ternper-

a t u r e below t h e previous set point. A few minutes later t h e h e a t pump comes back on-line and attempts to satisfy the demand. If it cannot, the furnace repeats the cycle.

One o t h e r method is to employ a hydronic system: w i t h a water-to- air heat exchanger in the supply duct downstream from the heat pump

condenser coil. T h i s would allow t h e h e a t pump to operate continuously

below the balance p o i n t . The hydronic supplemental h e a t may be provided by direct combustion o f f o s s i l f u e l , limiting the fossil fuel t o that

r e q u i r e d t o make up the difference between building l o a d and heat pump

capacity.

If t h e cooling function

of

the heat pump is not required o r desired,

the

indoor

c o i l

could be located upstream of the furnace. This would allow continuous operation af the heat pump below t h e

balance paint. Current practice r e f l e c t s the concern t h a t , i n t h e

c o a l i n g mode,the heat pump dehumidifies t h e r e t u r n air from the

conditioned space and any condensate e n t ~ a i n e d in t h e a i r stream could conrribute to corrosion

in

the heat exchanger o f t h e furnace.

ALTERNATIVE HEAT SOURCES FOR THE HEAT PIMP

Sub surface Ground

Throughout the l a t e 1 9 4 0 ' s and 5 0 t s , several electric utilities,

u n i v e r s i t i e s and other research organizations spent considerable effort

and money assessing t h e potential of the ground as a heat source and sink f o r the heat pump .5 The concensus of opinion at

the

time was t h a t

an a i r source heat pump would provide inadequate capacity under c o l d

ambient conditions. By the l a t e 1950fs, however, t h e work was d i s -

continued for a number of reasons some of which will

be

outlined here:

The air, as a heat source, i s more universally

p r e d i c t a b l e . A manufacturer can develop a product

l i n e which can be used over a wide range of- climatic

conditions. However, in the case of a ground source, soil properties t e n d to vary considerably from one region t o a n o t h e r , requiring sn site s o i l analysis for

each installation. I n particular, the amount and

configuration af t h e ground heat exchanger depends on the depth o f the water t a b l e

in

t h e ground. 2, W e r e summer c o o l i n g is of p ~ i m a r y importance the

ground is far from an i d e a l heat sink. The heat dissipated from t h e c o n d i t i o n e d space to t h e s o i l

drives t h e s o i l moisture from the v i c i n i t y of the

a c t i o n around t h e heat exchange surface, result

i n

a dramatic reduction

in

heat sink capacity.

3. The c o s t of installing t h e heat

transfer

c o i l was

perhaps the most important reason f o r

the

l i m i t e d

application of the ground

source

heat pump.

I n

some

i n s t a n c e s , however, depending on soil type, moisture

content or p r o x i m i t y of ground water, the ground

source heat pump may offer performance advantages over the air source heat pump, which might offset the

g r e a t e r capital investment required for the ground

coil.

h e o f t h e most d e f i n i t i v e

studies

was conducted in Port Credit,

Ontario between 1949-52,"is was a j o i n t study undertaken by

Ontario Hydro and t h e University of Toronto, An Ontario Hydro

employee's home was used f o r the s t u d y . Three separate ground c o i l s

were laid,

each w i t h an equivalent length

of

94.2 m (309 ft), at a

d e p t h of 1.5 m (5 ft),

on a

lot 30.5 by 45.7 m (100 by 1510 ft] [Figure 7 ) . The s o i l

on t h e

s i t e was described as a very f i n e sand w i t h a shallow water table. An e t h y l alcohol and water a n t i - f r e e z e

s o l u t i o n was circulated in the ground c o i l system which was

i n t e r -

faced with a water-to-air heat pump. The performance of the installation was monitored

for

t h r e e winters (Table 3 ) .

Because t h r e e separate ground c o i l s were available, the effect

of varying t h e installed ground c o i l capacity could be examined.

For t h e w i n t e r of 1950-51 (Figure 8) all t h r e e c o i l s were used and 4 8 per cent of the annual heating energy requiremenr was supplied by t h e heat extracted from the ground. A 3 . 7 3

kW

(5 hp) motor was

employed t o d r i v e t h e compressor. In t h e w i n t e r o f 1951-52 (Figure 9) o n l y one graund c o i l was employed and y e t 43 per cent of t h e annual h e a t i n g requirement w a s s u p p l i e d from the ground. The 3 . 7 3

kW

(5 hp)

compressor was replaced w i t h a 2.24 kW (3 hp] unit. In s p i t e of a 66 per c e n t

reduction

in

the e f f e c t i v e ground

coil

l e n g t h , t h e energy

savings were reduced by only 10 p e r cent. This i n d i c a t e s t h a t c o s t

optimization

studies a r e needed to select t h e most economical amount

o f ground c o i l and

size

of hear pump.

The Solar-Assisted Heat Puma

In t h e

p a s t , because of its intermittent nature and a need for a satisfactory means of collection, s o l a r energy was passed by in

favour of heat sources that were an i n d i r e c t result of solar input

(such as the ambient

air

and ground).

By

employing some form of

thcmal energy

storage,

the problems due to t h e intermittent nature

of solar radiation can be largely

overcome.

?'he efficiency of a flat p l a t e solar collector increases as the absorber temperature approaches t h e outdoor a i r temperature ( F i g u r e 1 0 ) .

To provide d i r e c t space heating, the collector would have to o p e r a t e

at a l a r g e temperature d i f f e r e n c e between absorber and outdoor

ambient u n d e r t y p i c a l midwinter Canadian conditions. However, by o p e ~ a t i n g t h e collector as t h e heat source f o r a h e a t pump, t h e collection temperature can be reduced, r e s u l t i n g in higher c o l l e c t o r efficiency and larger solar contribution o v e r t h e course of t h e heating season (Figure 1 0 ) .

Not unlike a conventional air-to-air heat pump, where t h e o u t d o o r evaporator surface area limits the "free" heat available to t h e c y c l e ,

t h e effectiveness o f t h e solas-assisted h e a t pump is governed by the callector area and thermal storage capacity. Increasing t h e s i z e of the c o l l e c t o r area ( f o r a f i x e d storage c a p a c i t y ) seems t o have a more

pronounced e f f e c t on system performance than increasing the size of s t o r a g e [ w i t h collector area fixed) .5 s 7 However, there a r e c o s t and

space limitations t h a t d i c t a t e t h e upper limit for collector area and

t h e r m a l storage volume, Far t h i s reason, a number of options should

be considered (see Figure 11

.

Provision could b e made f o r an outdoor

5

a i r - t o - a i r heat pump mode. '9 On m i l d , sunny days, t h e heat pump can

s a t i s f y t h e h e a t i n g demands o f t h e space u s i n g o u t d o o r air as a heat source, and the collector loop can charge t h e storage. Under mild,

overcast conditions, t h i s mode of operation will still satisfy the

h e a t i n g requirements, without depleting t h e thermal storage. A

system with heat storage on both evaporator and condenser s i d e s o f the _{5 7}

heat pump offers performance advantages over a s i n g l e starage system. 3 This would allow f o r heat pump o p e r a t i o n d u r i n g "off-peak" hours o n l y .

It would also be p o s s i b l e to operate with a smaller capacity heat pump,

running more or less continuously, not necessarily matching t h e

heating demands of the space at any given time. Further studies are

r e q u i r e d to assess t h e potential o f t h e s e and o t h e r s o l a r - a s s i s t e d

heat pump systems under Canadian c l i m a t i c conditions.

"ICE-MAKER1' HEAT PUMP

The most cornonly employed medium for thermal energy storage is water.

For each kilogram (pound) o f water, and C deg (F deg) d i f f e r e n c e between the upper and lower temperature limits of storage, 4 . 1 8 7 kJ (1 E t u ) o f heat can be stored. d thermal storage of 1.055 G J (10"tu] w i t h a useful temperature range of 27.7 C deg (50 F deg) requires 9072 kg

(20 000 l h ) o f w a t e T (9077

R;

2000 g a l . ) and would occupy a space of 9.07 m 3 (320 cu ft). T h i s would apply where only sensible h e a t of water

could be utilized, as in a solar h e a t i n g application, for example.

Consider instead, operation between -1 to IO"C (30 to 5 0 ' ~ ) ~ which encompasses the f r e e z i n g p o i n t of water. The heat capacity would be

379 kJJkg 1163 BtuJlb), the latent heat of f u s i o n of water yielding 335 k J / k g (144 Btu/lb) w i t h an additional 44 k ~ / k g (19 Btu/lb) s e n s i b l e heat (ice has only h a l f the s p e c i f i c heat o f water). For this temper- a t u r e range, a thermal storage of 1.055 GJ

[lo6

Btu] would r e q u i r e approximately 98

R

(613 gal.) of water that would occupy a space of 3 m 3 (107 cu ft) when frozen.

One such system b e i n g developed a t t h e

Oak

Ridge National

~ a b o r a t o r ~ ' is an "ice-maker" heat pump ( F i g u r e 3 2 ) .

An

aluminum p l a t e evaporator r e p l a c e s the conventional cross-flow a i r - t o - r e f r i g e r a n t

heat exchanger of a typical air-to-air heat pump. A circulating pump

delivers chilled water bo a distributor located above t h e evaporator p l a t e . The water f r e e z e s on t h e surface of the evaparator, which is maintained at approximately -7Oc ( 2 0 " F ) , liberating the latent heat o f fusion to the evaporating refrigerant. In time, the i c e b u i l d s up and

owing to its insulating effect performance, t e n d s to drop o f f . The i c e

is released from the evaporator p l a t e by using warm liquid refrigerant that leaves the condenses to temporarily raise the p l a t e temperature

above C'O (32*F). The

i c e

subsequently melts during mild weather.

I n

more northern regions, some t y p e o f

solar

collector may be required to

melt t h e ice and provide an auxiliary source of heat for the heat pump t o reduce the s i z e of the i c e stosage b i n .

SUMMARY

AND

CONCLUSIONS

1. In many cases today, the air-to-air heat pump is a b l e to

compete economically with electrical resistance heating systems. From a "prime" energy u s e point o f view, it

can raise the o v e r - a l l efficiency of elect~ical heating

from 28 to 42 per cent (Figure 1 3 ) . If the seasonal

performance factor could be improved,

in

t h i s case by

35 per c e n t , t h e a i r - t o - a i r heat pump would be mere resource energy efficient than conventional o i l and gas furnaces.

2. A n u m b e r o f a l t e r n a t i v e h e a t p u m p systems, l i s t e d b e l o w ,

require f u r t h e r investigation to assess their potential

for

residential

applicazion:

(i) ground source heat pump,

(ii) solar-assisted heat pump,

[iii] ''ice-maker" heat pump.

3 . There is little doubt as t o t h e performance improvements p o s s i b l e through t h e use of a l t e r n a t i v e heat sources f o r t h e heat pump but three important factors need t o be considered:

(i) i n i t i a l cost to the consumer, (i5) system r e l i a b i l i t y ,

(iii] load f a c t o ~ and t i m e of peak demand on the e l e c t r i c a l utility.

REFERENCES

1 Household Facilities and Equipment, Statistics Canada Catalogue

64-202

Annual,

May 1976.

2 Gardian Associates, Inc. Evaluation of t h e Air-to-Air Heat Pump for Residential Space Conditioning Prepared f o r Federal Energy

Administration, U . S . A . PB-255 652, April 1976, U.S. Dept. of

Commerce,

N.T.I.S.

268 p .

3 New Add-an Heat Pumps Works w i t h Furnace. A T s C o n d i t i o n i n g ,

Heating

E

Refrigeration N e w s , May 1977.

4 Solar E n e ~ g y Heat Pump Systems for Heating and C o o l i n g Buildings. Proceedings of Workshop a t Penn. S l a t e U. June 12-14, 1975.

Cosponsored by E . R . D . A . , Penn.State

U.,

and ASHRAE.

5 Joint

AEIC-EEI

Heat Pump Committee. Research Results Concerning

E a r t h as a Heat Source o r S i n k . Edison Electric Institute Bulletin,

S e p t . 1953, p . 355-358.

6 West, G.H. Residential Earth-Source Heat Pump. Ontario Hydro Research News, Vol. 11, No. 4 , Oct.-Dec. 1959, p . 1-3.

7 Abbaspour, M., Glicksman,

L.R.

The P r o p e r Use of Thermal Storages for a Solar Assisted Heat Pump Heating System,

A.S.M.E.

76-\'lA/HT- 7 6 .

8 Fischer, H.C. Thermal Storage Applications o f the Ice Maker

TABLE 2

Comparative C o s t s of Air-to-Air Heat

Pump

versus Resistance Heating

C o s t s Heat Pump Electric Furnace

7 kW (2 ton) (15 kW) Installed C o s t $2500 $1000

Difference

I n Installed Cost Maintenance Cost $50 $20 Difference

in Maintenance

Cost

Energy Cost

in F i r s t

Year $319

(@ 2.3$/kW-h) (13 870 kW-h)

TABLE 3

l+cat-pump- Syst em Performance Data for the Port Credit House

Heating Period

1949-50 1950-51 3951-52 Oct. 2h-June 9 a c t . 22-June 1 S e p t . 17-June 6

Degree Days, DC ( O F ) 3 816 (6 869) 5 634 (6 542) 3 993 ( 7 187)

Total Heat Energy Supplied, 14 910 14 883 13 671

k W q h (Btu) , from Ground Coi 1 (50 890 000) (5n 795 000) (46 659 000)

Total Electrical Energy Supplied For Heating, kW.h

To Compressor Motor

To Auxiliary Motors

To Resistance Heaters

Annual Coefficient of Performance

(COP) Over-a1 l

Heat Pump {Ground Coil and Compressor)

Heat Pump and Resistance Heaters Heat Pump and Auxiliary Motors

E f f e c t i v e Length of Ground C o i l , m Zft3

Total ILeat i n g Season, h

Total Operation o f Ileat hq, h

Annual Heating Requirements Supplied, %

By H e a t Pump

By Ground C o i 1

B y Rcs istance Heaters By Auxiliary Motors

- - G A S

-

C I U U I D E V A P O R A T O F ! co N U F N S E R L O W P R E S S U R E E X P A N S I O N H I G H P R E S S U R E V A L V E F I G U R E 1 B A S l C V A P O R C O M P R E S S B O N REF!? l G E R A T I O W C Y C L E

NON-RETWIN VALVE NON4ETURN VAPVE

F I G U R E 2

F I G U R E 3 P E R F O R M A N C E C H A R A C T E R 1 S T l C S OF A N A I R-TO-A I R H E A T PUMP 10 _{I I} H E A T PUMP C A P A C l T Y ( N O M I N A L 2-TON O R J - k W I NET B U I L D I N G H E A T L O S S 2

::

P O I N T O U T S I D E T E M P E R A T U R E H E A T P U M P C A P A C I T Y A N D B U I L D I N G H E A T 105s F O R A HOME I N THE O T T A W A A R E A

O U T S 1 DE TEMPERATURE F R E Q U E N C Y D l S T R l B U T l O N OF H O U R L Y T E M P E R A T U R E S I N O T T A W A B E L O W 6 5 ° F (18'C) A N N U A L H E A T I N G E N E R G Y A V A I L A B L E FROM H E A T O U T S I D E T E M P E R A T U R E A N N U A L H E A T I N G E N E R G Y R E Q U I R E D A S A F U N C T I O N OF O U T S I D E T E M P E R A T U R E A N N U A L E N E R G Y I N P U T T O H E A T PUMP AS h FUNCTION OF O U T S I D E T E M P E R A T U R E - 2 0 1 0 20 40 60

1

8 13 (-29) 1 ( - 1 8 1 1-71 I 4 1 1161

I

( 2 7 ) 3 O U T S I D E T E M P E R A T U R E

I

S T R E E T LINE,

-

N O T E . w A l l C o i l s 5 ' ( 1 . 5 m ) w B e l o w G r a d e

-

_- 0. N F I G U R E 7

S I T E P L A N OF GROUND S O U R C E HEAT PUMP k N S U E A T l O N

A T P O R T C A E D i T

A UNDISTURBED SOIL T E M P A T C O I L D E P T H B 3 N S f A N T A N F O U S H E A T W I T H D R A W A L i f A T E C A V E R A G E H I A T W I T H D R A W A L R A K E D C O I L T E M P F R A T U R E f l G U ' R E 8 G R O U N D C O I L PERFORMANCE - W I N T E R 1950 - 5 1 F I G U R E 9

S O L A R I R R A O I A T l O N = 2 5 0 Bkv/h f r 2 ( 7 8 8 W / m ( P L A I N A B S O R B E R , 1 T R A N S P A R F N T C O V E R ) 6 0

t

\

'

C O L L E C T I O N T E M P E R A T U R E \ REDUCED \ -TYPICAL- MIDWINTER IMPROVED I C O l l E C T l O M 3 0 ' \ I

I

E F F t t l E N C Y DIRECT S O L A R I - -

'

H E A T I N G

I

7 1 \ 0 " " I ' l " I ' I I ~ ~ ~ 1 ~ I I I I ~ \ *. F I G U R E 10 C O L L E C T O R PERFORMANCE W I T H A N D W I T H O U T H E A T P U M P

C O L L E C T O R

.

H! A T F l l ' A P F P A C E V a L T l N G 1 _{C I S C U I T} SFO?AGt A I R - T O - R F F A I G E R A N i C O L L E C T O R E V A P O R A T O R ~ o s 5 1 a l t C O N F ~ G U R A T ! O N S O F S O L A R - A S S I S T E D H E A T P U M o S HARVE5IING G A S VALVE

0 . 2 8 1 . a E L F C T R I C I T Y 0 . 4 2 H E A T PUMP S B F = 1 . 5 A T M O S P H E R E N A T U R A L G A S 0 . 6 0 5 CENTRAL REFJNlNG

PRCX*KII ION TRANSPORTATION (SPF=0.65) \

CENTRAL

TRANSPORTATION DISTRIBUTION

FROM UTILITY FURNACE H E A T I N G

PRCO1ICTIrn BY

PIPELINE (SPF = 0.7)

0 . 0 1

W i l l W a r y D e p e n d i n g On P i p s l i n e D i r t a n c c . C a p o c i t y , E t c .