Business machines
In offices, the presence of personal computers with peripheral devices and the commonplace use of other electrically energised equipment, gives a significant contribution to the sensible heat gains. Although power dissipations as high as 80 W m~2 have been quoted for dealers’ rooms in applications such as merchant bankers and stockbrokers offices, such figures are exceptional and are not representative of small power in ordinary offices. The actual rate of energy dissipation is generally less than the nameplate power on the items of equipment. Furthermore, machines are used intermittently and the power consumed when idling is less than when operating at full duty. The CIBSE Applications Manual AM7 (1992) gives useful details of suitable allowances for design purposes and quotes halfhour average powers in relation to nameplate powers.
Examples of some values are given in Table 7.17, based on an office total floor area of 583.2 m2 (see Figure 7.21) and using data from CIBSE AM7 (1992).
Table 7.17 Examples of powers for business machines. (For an office floor area of 583.2 m2)

18 modules x 2.4 m = 43.2 m
Fig. 7.21 A notional office plan area, used to establish the typical power dissipation from office Machines (Table 7.17). 
From the above table the specific totals, per unit of floor area, are 52.1 W m“2 for nameplate powers and 18.5 W m2 for halfhour averages. Note that, for the above example, the coffee machines make a very large contribution to the total nameplate power but a good deal less to the halfhour average. For other cases the figures would be different.
Diversity factors that may be applied to the nameplate power to represent the load on the air conditioning system for the whole building are difficult to establish with any certainty. From Table 7.17 the ratio of total halfhour average power to total nameplate power gives a value of 10790/30375 = 0.36. From CIBSE AM7( 1992) such ratios can be deduced and some are given in Table 7.18.
The variation in the values quoted depends on the assumptions made for the time that the machines are running or idling.
Table 7.18 Ratios of halfhour average powers to nameplate powers according to CIBSEAM7 (1992)

Reproduced by kind permission of the CIBSE, from AM7 (1992). See also CIBSE Guide A2 (1999): Internal heat gains, 6.4. 
A cautious conclusion is that a typical allowance for small power for business machines is from 10 W m2 to 20 W m2, in the absence of firmer information to the contrary. As regards a diversity factor to be applied when calculating the total maximum sensible heat gain for the whole building (in order to determine the size of refrigeration plant) a value of 0.65 to 0.70 seems reasonable. For the purpose of calculating the maximum sensible heat gain for a particular room or module (to determine the amount of supply air necessary or the size of air conditioning unit to select), a diversity factor of unity must be applied to the heat gain from business machines, people and electric lights.
EXAMPLE 7.18
The dimensions of a westfacing module in a lightweight building (150 kg m2) are: 2.4 m width x 2.6 m floortoceiling height x 6.0 m depth. The floortofloor height is 3.3 m. A double glazed, 6 mm clear glass window of 2.184 m2 area is in the only exterior wall. Rooms on the other five sides of the module are air conditioned to the same temperature. The building is at latitude 51.7°N, approximately, and the window is protected by internal Venetian blinds, assumed to be drawn by the occupants when the window is in direct sunlight. The [/value of the window is 3.0 W m2 K_1. The time lag of the wall is 5 h and its decrement factor is 0.65.
Outside state: 28°C drybulb, 19.5°C wetbulb (sling), 10.65 g kg1.
Room state: 22°C drybulb, 50 per cent saturation, 8.366 g kg1.
(a) Making use of Tables 7.8, 7.9 and 7.10 calculate the sensible and latent heat gains at 1500 h suntime in July. When calculating the heat gain through the wall use the floortofloor height for determining the wall area.
(b) Repeat the calculation of the sensible heat gain using the data from Tables 7.11 and 7.12 for the solar load through glass. Take a haze factor of 0.95 and a blind factor of 0.54 (Table 7.6).
(c) Repeat the calculation for the sensible heat gain using the data from Tables 7.13 and 7.14 for the solar load through glass.
(d) Repeat the calculation using the solar heat gain data in Table A9.15 in the CIBSE Guide A9 (1986).
(e) Compare the results.
Answer
(a) First the relevant solair temperatures must be established from Table 7.8 for the month of July and any necessary corrections made:
Design outside air temperature at 15.00 h sun time: 28°C
Tabulated outside air temperature at 15.00 h sun time: 25.3°C
Correction to be applied to tabulated value: +2.7 K
Time of heat gain to room: 15.00 h
Time lag of wall: 5 h
Time of relevant solair temperature 10.00 h
Tabulated solair temperature at 10.00 h 25.0°C
Correction: +2.7 K
Corrected solair temperature at 10.00 h 27.7°C
Tabulated 24 h mean solair temperature: 23.0°C
Correction: +2.7 K
Corrected 24 h mean solair temperature 25.7°C
Use the simplified equation (7.24) for the calculation of heat gain through the wall.
From Tables 7.9 and 7.10 the following is determined for a lightweight building with blinds closed (because the window is in direct sunlight), 12 h plant operation and air temperature control.
Solar air conditioning load through a west window at 15.00 h sun time in July: 205 WnT2. Correction factor for a double glazed, clear, plate glass window: 1.00.
Sensible heat gain calculation:
TOC o "15" h z Glass transmission (equation (7.17)): W
2.184 x 3.0 x (2822)= 39
Wall (equation (7.24)):
(3.3 x 2.4 — 2.184) x 0.45R25.7 — 22) + 0.65(27.7 — 25.7)] = 13
Infiltration (equation (7.41)):
0.33 x 0.5 x (2.4 x 2.6 x 6.0)(28 — 22) = 37
Solar gain through glass:
2.184x 1.00×205 = 448
People: 2 x 90 = 180
Lights: 17 x 2.4 x 6.0 = 245
Business machines: 20 x 2.4 x 6.0 = 288
Sensible heat gain = 1250
Latent heat gains:
34 W 100 W 134 W 
Infiltration by equation (7.42):
0. 8 x 0.5 x (2.4 x 2.6 x 6.0) x (10.65 — 8.366) = People: 2 x 50 =
Latent heat gain =
(b) From Table 7.11: maximum solar intensity: 695 W From Table 7.12: storage load factor: 0.67 From Table 7.6 the shading coefficient: 0.54 Solar load: 0.95 x 0.54 x 0.67 x 695 = 239 W nT2 Total sensible heat gain from (a): 1250 W
802 W 522 W 1324 W 
Solar gain through glass from (a): 448 W
Remaining other sensible gain: Solar gain: 2.184 x 239 =
Sensible heat gain:
(c)
,2 
From Table 7.13:
Solar cooling load (271 + 325)/2 = 298 W m“
From Table 7.14, the correction factor for shading is 0.95 and the airpoint control factor (because the tabulated data are for the control of dry resultant temperature) is 0.91 Remaining other sensible gain: 802 W
1365 W 
Solar gain: 2.184 x 298 x 0.95 x 0.91 563 W
Sensible gain:
(d) From CIBSE Table A9.15 (1986) the cooling load due to solar gain through glass is 270 W m2, the shading factor is 0.74 and the airpoint control factor is 0.91.
Remaining other sensible gain: 802 W
397 W 1199 W 
Solar gain: 2.184 x 270 x 0.74 x 0.91 Sensible gain:
(e) Comparing the results:
Sensible 
Glass 

Gain 
Solar gain 

(d) Table A9.15 (CIBSE 1986) 
1199 W 
100% 
397 W 
100% 
(a) Tables 7.9 and 7.10 (Haden Young) 
1250 W 
104% 
448 W 
113% 
(b) Tables 7.11 and 7.12 (Carrier Air 

Conditioning Company 1965) 
1324 W 
110% 
522 W 
131% 
(c) Tables 7.13 and 7.14 (CIBSE A2 1999) 
1365 W 
113% 
565 W 
142% 
Note that the approximate percentages of the components of the sensible gain to a typical office module are: glass transmission 3 per cent, wall transmission 1 per cent, infiltration 3 per cent, solar gain through glass 36 per cent, people 14 per cent, lights 20 per cent and business machines 23 per cent. It is also worth noting that the approximate specific sensible gain is about 80 to 90 W m2, referred to the treated modular floor area.
The effects of changing the summer outside design state and the state maintained in the room are worth consideration. Tables 5.3 and 5.4 give details of the percentages of the hours in the summer that outside drybulbs and wetbulbs were exceeded at Heathrow for the period 197695. If lower outside design temperatures are chosen the heat gains will be reduced but the chances of the air conditioning system not giving full satisfaction are increased. Similarly, raising the drybulb temperature maintained in the room, with the risk of less comfort, may be a consideration. Further, if the outside wetbulb or the room enthalpy are altered there will be an effect on the fresh air load (see sections 8.1 and 8.2). All these matters have implications for comfort, the satisfaction of the client, capital costs and running costs. The benefits of any measure considered must be balanced against its disadvantages, by considering Table 5.3 or 5.4.
EXAMPLE 7.19
{a) Ignoring heat gains through the wall, which are trivially small, investigate the effect on the sensible heat gain of changing the outside and inside summer design states for the module used in example 7.18. Use Tables 7.9 and 7.10 for calculating the solar heat gains through glass.
Answer
The sensible heat gains from solar gain through glass, lights and business machines are unchanged at: 448 + 245 + 288 = 981 W. The sensible gain from people reduces and the latent gain increases as the room temperature rises (see Table 7.16). The sensible gain from infiltration reduces for a given room condition, as the outside drybulb goes down. The sensible heat gains are considered as follows.
Sensible heat gain: 

O 00 CN 
27° 
26‘ 

Tr 
22° 
22° 
22‘ 
Glass 
39 
33 
26 
Infiltration 
37 
31 
25 
People 
180 
180 
180 
S + L + M 
981 
981 
981 
Total 
1237 
1225 
1212 
If we assume that the fresh air allowance is 1.4 litres s’1 m2 over the modular floor area of 14.4 m2 the supply is 20.16 litres s1 of fresh air with a specific volume, v, expressed at the room state. Assuming the room is at 50 per cent saturation, the latent heat gain by infiltration can be calculated by equation (7.42) and we can also calculate the fresh air load (see example 8.1) by means of equation (7.43):
Qfa = (vfv^K — hr))
Where <2fa is the fresh air load in kW, vt is the volumetric flow rate of fresh air at temperature
T in m3 s~ vt is the specific volume of the fresh air at temperature t, h0 is the enthalpy of the outside air in kJ kg1 and hr is the enthalpy of the room air in kJ kg“1.
The following table is compiled.
To (°C) 
28 
27 
27 
26 
26 
To (°C) 
20 
19 
18 
19 
18 
Tr (°C) 
22 
22 
22 
22 
22 
8o (g kg“1) 
10.65 
9.719 
8.354 
10.18 
8.859 
Gr (g kg“1) 
8.366 
8.366 
8.366 
8.366 
8.366 
People latent gain, 
100 
100 
100 
100 
100 
See Table 7.16 (W) 

Infiltration latent gain, 
34 
20 
0 
27 
7 
See equation (7.42) (W) 

People + infiltration 
134 
120 
100 
127 
107 
Latent gain (W) 

H0 (kJ kg“1) 
55.36 
51.96 
48.47 
52.10 
48.74 
Hr (kJ kg“1) 
43.39 
43.39 
43.39 
43.39 
43.39 
Vr (m3 kg1) 
0.847 
0.847 
0.847 
0.846 
0.847 
Fresh air load (see 
285 
204 
121 
207 
127 
Equation (7.43) and section 6.7) (W) 

Sensible gain 
1636 
1537 
1446 
1530 
1442 
+ latent gain + fresh air load (W) 
Interpreting the data in Table 5.3 is difficult. One way might be to use midvalues of the bands and the averages of the limits of the bands of drybulb and wetbulb temperatures quoted:
For 28°C drybulb and 20°C wetbulb the average is (0.77 + 0.32 + 0.31 + 0.25)/4 = 0.41 per cent of the four summer months, namely 12.0 hours. For 27°C drybulb and 19°C wet — bulb it is 0.77 per cent or 22.5 hours. For 27°C drybulb, 18°C wetbulb it is (0.42 + 0.77)/
2 = 0.60 per cent or 17.6 hours. For 26°C drybulb and 19°C wetbulb it is (1.32 + 0.77)/
2 = 1.04 per cent or 30.4 hours and for 26°C drybulb and 18°C wetbulb it is (1.37 + 1.32 + 0.42 + 0.77)/4 = 0.97 per cent which is 28.4 hours. The results are summarised in the following table.
1. Drybulb 
28° 
27° 
27° 
26° 
26° 
2. Wetbulb 
20° 
19° 
18° 
19° 
18° 
3. Per cent exceeded 
0.41% 
0.77% 
0.60% 
1.04% 
0.97% 
4. Hours exceeded 
12.0 h 
22.5 h 
17.6 h 
30.4 h 
28.4 h 
5. Sensible gain 
100% 
99% 
99% 
98% 
98% 
6. Latent gain 
100% 
95% 
83% 
97% 
90% 
7. Fresh air load 
100% 
72% 
43% 
73% 
45% 
8. Totals 
1636 W 
1537 W 
1446 W 
1530 W 
1442 W 
100% 
94% 
88% 
94% 
88% 
208 Heat gains from solar and other sources Conclusions
1. The sensible gains do not change very much and the reduction is probably within the accuracy of their calculation. Any effect they have is most likely to be on the size of the air handling plant and duct system, although this depends very much on the type of system adopted.
2. For the case of an outside state of 27° drybulb, 18° wetbulb, the large drop in the latent gain to 83 per cent must be disregarded since this is entirely due to the reduction of the infiltration gain to zero, because the outside moisture content is less than the room moisture content. The allowance of half an air change per hour for infiltration is very much open to question and the reduction must be ignored.
The reduction of the latent gains to 95 per cent and 90 per cent for the cases of 21° dry — bulb, 19° wetbulb and 26 per cent drybulb, 18° wetbulb look interesting but the latent gains are a fairly small proportion of sensible plus latent gain, namely, about 10 per cent. Any benefit in reducing the capital or running cost is likely to be in the size of the refrigeration duty. The impact of the reduction in latent gains on this is smaller still, at about 7 per cent.
3. If, as an approximation, the cooling load is taken as the sum of the sensible gain, the latent gain and the fresh air load, we can see that this is probably the most significant factor. There is an established correlation between the cooling load and capital cost. The running cost is also likely to correlate although other factors, such as system choice and the quality of maintenance, intervene. It must be remembered that this example is based on the heat gains and cooling load for a single, westfacing module and diversity factors for people, lights, and business machines have not been considered, whereas they would be for an entire building. Neither has the natural diversity in the solar load through glazing, as the sun moves round the building during the day, been taken into account. Nevertheless, designing the system for an outside state of 27°C drybulb with 19°C wetbulb (screen) gives a 6 per cent reduction in the cooling load (implying a 6 per cent reduction in capital cost). This is at the expense of a failure in apparant comfort satisfaction for 22.5 h in the four summer months. The decision is commercial and the client must be fully aware of the risks involved for the benefit obtained.
The view of the author is that, with global warming likely to become increasingly important over the next 20 years, prudence is to be recommended. Air conditioning systems are likely to be required to do more than they were initially designed for in the immediate future.
EXAMPLE 7.20
Using the modular details from example 7.18 calculate the sensible and latent heat gains at 1500 h sun time on 21 January, assuming the location to be Perth, Western Australia. Take the same values for heat gains from lights and business machines, assume two people present and half an air change of infiltration. The window is single glazed (U = 5.6 W mf2 K1) and fitted with internal Venetian blinds. The latitude of Perth is 32°S. Make the following design assumptions:
Outside state: 35°C drybulb, 19.2°C wetbulb (screen), 5.876 g kg’1, 50.29 kJ kg1,
0. 8809 m3 kg’1.
Room state: 25°C drybulb, 40 per cent saturation, 8.063 g kg1, 45.69 kJ kg1.
The load due to sensible heat gain through single glass by solar radiation for a fast response building with internal blinds, used intermittently when the sun shines on the west face of the building, is 241 W m1, based on data from the CIBSE Guide A2 (1999). This should be increased by 7 per cent because the earthsun distance is 3.5 per cent less in December than in June and the intensity of solar radiation follows an inverse square law with respect to distance. See Table 7.4 and section 7.16.
Answer
First the relevant solair temperatures should be calculated. Solair temperature is given by equation (7.21). Using the results of example 7.4, equation (7.9) and the data in Tables 7.3 and 7.4, the calculation of the solair temperature at a particular time is possible, but calculating the heat gain by equation (7.24) requires a knowledge of the 24 hour mean sol — air temperature. This is not possible here, without detailed weather data, on an hourly basis, for January at Perth. However, the heat gain through a wall usually amounts to about
1 per cent of the whole (see example 7.18) and is ignored in this case.
TOC o "15" h z Sensible heat gain calculation: W
Glass: 2.184 x 5.6 x (35 25) 122
Infiltration: 0.33 x 0.5 x (2.4 x 2.6 x 6.0) x (35 — 25) 62
Solar through glass: 2.184 x 241 x 1.07 563
People*: 2 x 75 150
Lights: 17 x 2.4 x 6.0 245
Business machines: 20 x 2.4 x 6.0 288
Total sensible gain 1430
*Refer to Table 7.16
Specific sensible gain per unit of floor area:
1430/14.4 = 99.3 W m“2 Latent heat gain calculation:
TOC o "15" h z Infiltration (equation (7.42)): W
0. 8 x 0.5 x 2.4 x 2.6 x 6.0(5.876 — 8.063) 33
People:
2 x 65 130
Total latent gain 97
Since any rate of natural infiltration is uncertain, it might be prudent to take the latent gain as 130 W. However, local custom should be considered as well as meteorological data (see section 5.11). There is an onshore breeze from the Indian Ocean that often occurs in the afternoons during summer months in Perth, which reduces the drybulb temperature and increases the moisture content. This could make the assumed outside design state temporarily irrelevant. The contribution of any infiltration in the latent heat gain calculation would then be positive with the total latent gain exceeding 130 W.
1. (a) Why do the instantaneous heat gains occurring when solar thermal radiation passes
Through glass not constitute an immediate increase on the load of the airconditioning plant? Explain what sort of effect on the load such instantaneous gains are likely to have in the long run.
(b) A single glass window in a wall facing 30° west of south is 2.4 m wide and 1.5 m high. If it is fitted flush with the outside surface of the wall, calculate the instantaneous heat gain due to direct solar thermal radiation, using the following data:
Intensity of direct radiation on a plane normal to
TOC o "15" h z the sun’s rays 790 W m2
Azimuth of the sun 70° west of south
Transmissivity of glass 0.8
Answers (,b) 870 W.
2. A window 2.4 m long x 1.5 m high is recessed 300 mm from the outer surface of a wall facing 10° west of south. Using the following data, determine the temperature of the glass in sun and shade and hence the instantaneous heat gain through the window.
TOC o "15" h z Altitude of sun 60°
Azimuth of sun 20° east of south
Intensity of sun’s rays 790 W m2
Sky radiation normal to glass 110 W m2
Transmissivity of glass 0.6
Reflectivity of glass 0.1
Outside surface coefficient 23 W nT2 K1
Inside surface coefficient 10 W m2 K1
Outside air temperature 32°C
Inside air temperature 24°C
Answers
37.9°C, 29.6°C, 983 W.
3. An air conditioned room measures 3 m wide, 3 m high and 6 m deep. One of the two
3 m walls faces west and contains a single glazed window of size 1.5 m by 1.5 m. The window is shaded internally by Venetian blinds and is mounted flush with the external wall. There are no heat gains through the floor, ceiling, or walls other than that facing west and there is no infiltration. Calculate the sensible and latent heat gains which constitute a load on the air conditioning system at 16.00 h in June, given the following information.
Outside state Inside state Electric lighting 
28°C drybulb, 19.5°C wetbulb (sling) 22°C drybulb, 50% saturation 33 W per m2 of floor area
Number of occupants
Heat liberated by occupants
Solar heat gain through window with
258 W m~2 1.7 Wm2K‘ 5.7 W m“2 K"1 5 hours (= <)>) 0.62 (=/) 
Venetian blinds fully closed [/value of wall [/value of glass Time lag for wall Decrement factor for wall
Diurnal variations of air temperature and solair temperature are as follows:
09.0
Suntime Air temperature (°C) Solair temperature (°C) 
10.00 11.00 12.00 13.00 14.00 15.00 16.00
20.6 22.0 23.3 24.7 25.8 26.8 27.5 28.0
23.7 25.3 26.8 28.3 39.4 47.3 53.6 57.0
The mean solair temperature over 24 hours is 29.9°C (= fem).
The heat gain through a wall, q^), at any time (0 + ((>), is given by the equation:
9(0+0) — ( /cm — + UA(te — tem)f
Where te is the solair temperature t is the inside air temperature 0 is the time in hours.
Answers 1662 W, 200 W.
4. (a) Derive an expression for solair temperature.
(.b) Using your derived expression determine the solair temperature for a flat roof if the direct radiation, normal to the sun’s rays, is 893 W nT2 and the intensity of scattered radiation normal to the roof is 112 W m2 Take the absorption coefficient of the roof for direct and scattered radiation as 0.9, the heat transfer coefficient of the outside surface as
22.7 W m2, the outside air temperature as 28°C and the solar altitude as 60°C.
(c) Given that the time lag of the roof structure is zero and its decrement factor is unity, calculate the heat gain to the room beneath the roof referred to in part (b) if the [/value of the roof is 0.5 W nT2 K"1 and the room temperature is 22°C. The mean solair temperature over 24 hours is 37°C.
Answers
(b) 63.1°C, (c) 20.6 Wm“2.
5. Repeat the calculation of the sensible gain for the module used in example 7.18 but for an eastfacing module, with a wall [/value of 0.6 W m“2 KT1 and an infiltration rate of 2 air changes per hour.
Answer
1111 W.
Symbol A 
Unit Wm“2 M2 M2 M M2 M2 Medians Degrees Degrees 
Description
Apparent solar radiation in the absence of an atmosphere
Solar radiation constant area
Surface area of a structural element in a room internal duct dimension floor area
Area of glass or area of opening in a wall
Angular movement of the sun
A A’ B C C D D Fb F* 
Altitude of the sun
Altitude of the sun at noon
Atmospheric correction factor
M KJ kg’1 K1 m Degrees 
Internal duct dimension
Dimensionless constant
Specific heat capacity
Internal duct diameter
Declination of the sun
Shading factor
Airpoint control factor
Angle factor for the ground
Angle factor for the sky
Room conduction factor with respect to the air node
* ay / 80 gr h 
Room admittance factor with respect to the air node
G kg 1 g kg1 degrees W nT2 KT1 W nT2 KT1 W nT2 KT1 
Decrement factor
Outside air moisture content
Room air moisture content
Hour angle
Coefficient of heat transfer
Hsi 
Inside surface film coefficient of heat transfer
KJ kg kJ kg 
Outside surface film coefficient of heat transfer
Enthalpy of the outside air
Hr I 
Enthalpy of the room air
W m"2 or kW m~2 Wm’2 W nT2 
Intensity of direct solar radiation on a surface
Normal to the rays of the sun
Component of direct solar radiation normal to a horizontal surface
Intensity of radiation reflected from surrounding surfaces
W nT Wm 
2 
2 
Wm 
Intensity of diffuse (sky) radiation normally incident on a surface intensity of total radiation on a surface component of direct solar radiation normal to a vertical surface
H 
2 
Wm 
Component of direct solar radiation normal to a tilted surface
Angle of incidence of a ray on a surface latitude of a place on the surface of the earth duct length
Degrees Degrees M M M or mm degrees H1 M W nT2 W W m~2 KW W W W W W W 
I L I M N P Q & Qfa Qu Qm Qs Фsi (2e Фe+<t> ?max 4s R Rsi Rso Sc SHGC T 
Thickness of the insulation on the duct
Dimension of a hypotenuse formed by R and x
Wallsolar azimuth
Number of air changes per hour
External duct perimeter
Rate of heat flow
Heat transfer through a duct wall
Rate of heat entry to an outer wall/roof surface
Fresh air load
Wm Wm M or mm Wm“2 m2 K W“1 m2 K W1 H K °C H °C °C °c °c °c °c °c °c °c °c °c °c °c °c °c W m“2 KT1 3 M 
Latent heat gain by natural infiltration mean rate of heat flow through a wall or roof cooling load due to solar gain through glass sensible heat gain by natural infiltration rate of heat flow into a room at time 0 rate of heat flow into a room at time 0 + <j) maximum instantaneous sensible solar heat gain through glass
Specific cooling load due to solar gain through glass depth of a window recess
Remainder term to cover longwave radiation exchanges thermal resistance of an inside surface air film thermal resistance of an outside surface air film shading coefficient solar heat gain coefficient sun time
Absolute temperature drybulb temperature time
Wetbulb temperature
Air temperature
Solair temperature
Inside environmental temperature
Mean solair temperature over 24 hours
Mean glass temperature
Outside air drybulb temperature
Room air temperature
Ambient air drybulb temperature
Mean radiant temperature
Inside surface temperature
Lso ‘ Slll H H U V 
Outside surface temperature
Mean inside surface temperature
Initial air temperature in duct
Final air temperature in duct
Overall thermal transmittance coefficient
Volume of a room
V mean air velocity in a duct m s_1
Velocity of airflow through an opening m s1
Vt volumetric flow rate of fresh air at temperature t m3 s1
Vt specific volume of air at temperature t m3 kg1
X horizontal coordinate m or mm
Y admittance of a surface W m2 K~*
Y vertical coordinate m or mm
TOC o "15" h z dimensionless parameter related to duct heat gain —
Z azimuth of the sun degrees
A absorptivity of glass for direct solar radiation —
A’ absorptivity of glass for scattered solar radiation —
Ap pressure drop through an opening Pa
At temperature change per metre of duct length K trf1
8 angle between the ground and the horizontal degrees
Angle of a surface with the horizontal degrees
TOC o "15" h z 0 time h
X thermal conductivity of the insulation on a duct W m1 K1
P density of air kg itT3
Glass reflection coefficient for direct solar radiation —
Reflectivity of the ground —
X glass transmission coefficient for direct radiation —
X’ glass transmission coefficient for sky radiation —
<)) time lag h
Posted in Engineering Fifth Edition