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Hot Water Heating
Module HWC
Hot Water Heating
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Hot Water Heating
CONTENTS SCOPE AND OBJECTIVES.............................................................................4 THE NEED FOR A HEATING SYSTEM ..........................................................5 FUNDAMENTALS............................................................................................6
Temperature.................................................................................................6 Heat..............................................................................................................7 Specific heat capacity...................................................................................7 Heat flowrate or heat transfer rate................................................................7 Transfer of heat ............................................................................................8
THERMAL ENVIRONMENT ............................................................................9 Heat losses.................................................................................................10 Design temperatures ..................................................................................14
BASIC ELEMENTS OF A HOT WATER HEATING SYSTEM........................20 Feed and expansion...................................................................................20 Air removal .................................................................................................23 Standard Symbols and Abbreviations ........................................................25
HEAT EMITTERS ..........................................................................................26 Natural convective......................................................................................26 Forced convective ......................................................................................26 Radiant .......................................................................................................26 Emission figures .........................................................................................27
PIPING SYSTEMS.........................................................................................28 Gravity systems..........................................................................................28 Pumped systems........................................................................................28 Flows of liquids through pipes ....................................................................32
PIPE SIZING..................................................................................................39 Introduction.................................................................................................39 Step-By-Step Pipe Sizing.........................................................................39
SELECTING THE BOILER AND PUMP ........................................................50 Boiler heat capacity ....................................................................................50 Varying the flow or return temperature .......................................................51 Pumps and pump sizing .............................................................................54 Pump characteristics ..................................................................................55
SYSTEM CHARACTERISTICS .....................................................................59 The system curve .......................................................................................59 The pump curve .........................................................................................60
POSITION OF PUMP IN SYSTEM ................................................................63 Introduction.................................................................................................63 Summary of pump positions .......................................................................69 Inverter Driven Pumps................................................................................71 Balancing the system .................................................................................72
AUTOMATIC CONTROL OF A HEATING SYSTEM .....................................76 Boiler control ..............................................................................................76
TEMPERATURE CONTROL .........................................................................78 Room thermostat........................................................................................78 Individual radiator control ...........................................................................78 Mixing valves with compensator (i.e. Variable Temperature (‘VT’) Circuits)...................................................................................................................79
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Diverting valves ..........................................................................................80 Zone control ...............................................................................................81 Time control................................................................................................82 Frost protection ..........................................................................................83 Building Management Systems (BMS) or Building Energy Management Systems (BEMS) ........................................................................................83 Sizing the control valve ..............................................................................83 Valve authority............................................................................................86 The need for balancing three-way valves...................................................88 Effects of two-port valves ...........................................................................90 Multi-zone circuits with three-way control ...................................................91 Automatic balancing ...................................................................................92
HIGH TEMPERATURE HOT WATER SYSTEMS .........................................94 DOMESTIC HOT WATER SUPPLIES ...........................................................96
DHW systems.............................................................................................96 HWS storage and boiler power...................................................................97 Secondary piping........................................................................................97 Calculation of hot water demand................................................................98 BS6700 Loading Units................................................................................99 Secondary Circulation Pumps ..................................................................103
LEGIONELLOSIS ........................................................................................104 Background ..............................................................................................104 Regulation and Enforcement ....................................................................105 Maintenance Precautions and Control Measures.....................................106
SUMMARY...................................................................................................108
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Hot Water Heating
SCOPE AND OBJECTIVES
This module is not intended to turn anyone into a fully-fledged heating designer or building services practitioner. There are many learned textbooks around that are far better suited to help in that task. Indeed, heating, or perhaps more correctly these days, environmental engineering is an enormous science covering many more subjects than simply hot water heating systems. University and other centres of excellence provide courses for in-depth studies of building services and environmental engineering.
The purpose of this module is rather to take a small (but important) part of the building service package - hot water heating and domestic hot water systems - and to look broadly at some fundamental aspects of their design and function. This is intended to provide some basic knowledge and help in this area for people such as the Engineering Manager or perhaps Plant Engineer or Hospital Engineer, who does not come from a building services background, but who does have to interface with such systems, whether occasionally or on a day-to-day basis.
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THE NEED FOR A HEATING SYSTEM Keeping warm is one of man’s most primitive instincts – borne from his caveman days and experimentation with fire. The Romans turned heating into a science over two thousand years ago with under-floor heating and other ingenious devices, and we have been copying and striving to improve upon their ideas ever since. The human body functions best in temperatures of between 16oC and 22oC. It was no accident that early civilisation developed and flourished around the 21oC isotherm. The more civilised we became the more sophisticated were the demands to control and improve the environment in which we live, work and play. Buildings may be warmed by the sun in those countries enjoying a hot climate; here the need is often to cool the environment. Even in temperate zones, given sufficient insulation, internal spaces may be warmed by the internal heat from machinery, lights, and even the occupants. But, in many climes, especially in the extreme northern and southern hemispheres, it is necessary to warm buildings by burning a fuel or consuming electricity, which in turn is derived from a fuel. It is to the means for performing these functions and distributing the heat efficiently to places where it is needed at the right time and in the correct amount, that the heating engineer directs his attention. Warmth alone is not the sole criterion of comfort. There must also be a pleasant atmosphere which may involve ventilation to remove stale air and replenish with fresh air; control of the moisture in the air (humidity) and perhaps complete air conditioning for cooling in summer as well as heating in winter. Fundamentally, all these and other related problems can be resolved simply into questions of how to add or remove heat, how to add or remove moisture, and how to circulate air. We can lump all this together and call the whole subject ‘comfort conditioning’. The water heating system comes under the heading of adding heat.
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Hot Water Heating
FUNDAMENTALS
Temperature The sensation of heat is relative; boiling water is at a higher temperature than warm water; cold water is at a higher temperature than ice. Heat will flow from a hot body to a cooler body, the rate of travel being proportional to the temperature difference. We have adopted ‘scales’ to measure temperature, and the instrument used for this purpose is a thermometer. The thermometer measures dry bulb air temperatures (Ta). If the thermometer is placed in a room in which the air, walls, floor and ceiling are all at the same temperature, assuming we are comfortable at this temperature, then the thermometer could be regarded as indicating a satisfactory index of comfort. Such conditions of uniformity of temperature of air and fabric surfaces rarely occur. For instance, in a warm room, heat will pass through a window at a greater rate than through the other surfaces of the room. In order to counteract this, the room air temperature would need to be raised or some form of radiant heat emitter provided. Under these conditions, where air temperature and surface temperatures of the enclosed space are significantly different, the ordinary thermometer fails as a true index of comfort. A better indicator of the mean room temperature is the globe thermometer. The globe thermometer (Fig. 1) comprises a thermometer inserted into a matt black sphere of copper or other thin material about 150 mm in diameter. When suspended in a room, the black sphere will absorb heat radiation or emit radiation to cold surfaces and achieve some kind of thermal balance, comparable to what is termed the mean radiant temperature (tr).
Loose fitting stopper
Ordinary thermometer
150 mm diameter blackened copper globe
Fig. 1 Globe thermometer
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Hot Water Heating
Heat Energy can move from one place to another in more than one way; one might be by thermal energy transfer (heat), another by mechanical energy transfer (work), and yet another by chemical interaction. This module is concerned with the processes of heat transfer, such as that associated with heat exchangers transferring heat from steam to water, or from water to air via low temperature hot water ‘radiators’. But first, let’s examine a little further the concept of heat. Heat is to thermal energy what work is to mechanical energy. All systems contain ‘internal energy’, proportional to molecular activity. When two systems with different levels of internal energy interact, energy can be transferred between them in the form of thermal energy and/or work energy. Heat is the quantity of thermal energy that might pass between the two systems under such circumstances. The amount of heat transferred can be calculated, but it should be remembered that thermometers only measure temperature and not the quantity of heat given out or received by a body. Heat flows by conduction, convection or radiation from one place to another owing to a temperature difference between the places; this concept obeys the second law of thermodynamics. The derived SI unit of heat is the joule (J) which is equivalent to the work done or the energy required to exert a force of one Newton through a distance of one metre; that is, 1 J = 1 N m. In base terms, the SI unit for energy is kg m2/s2.
Specific heat capacity The specific heat capacity of a body of any kind is defined as the heat required to raise its temperature by 1°C; the practical engineering unit for this is normally taken as kilojoules per kilogram per degree Celsius (kJ/kg°C), and is sometimes described as ‘specific heat’. If we take equal masses of water and oil and heat them separately for the same time with exactly the same heat source, the oil temperature might rise by 4°C in two minutes, whereas the water temperature might rise by only 2°C. Since the rate of heat supply is the same in both cases, oil has a smaller heat capacity than an equal mass of water. The comparison of the heat capacities of various substances is known as their specific heat capacity or, colloquially, specific heat.
Heat flowrate or heat transfer rate As intimated previously, heat is always on the move, it cannot exist of itself as a self-contained body; heat is continuously being transferred to or from any one body in the universe to or from its surroundings. So, in fact, the heat ‘contained’ by steam or water at a certain temperature is the balance of heat to and from its surroundings at any time. It is not possible, in the strict thermodynamic sense, to have a ‘reservoir’ of heat (in the same way as water in a bucket), although it rather suits us to think of the thermal energy contained in steam or water as having a ‘potential’ to transfer heat. In this way, we can consider water or steam as containing a certain amount of heat
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Hot Water Heating
that can be used to raise the temperature of another body surrounding it. This philosophy allows us to make practical heat transfer calculations. To satisfy these practical needs, if we consider that a body contains a certain amount of heat ready to give up to its surroundings, we can also consider how quickly it does so. The rate at which heat travels is usually of more importance to us than how much is available for transfer (although this is important too). If the unit of heat is the joule, the rate at which heat travels is measured in joules per second (J/s); another name for 1 J/s is the watt (W). The term ‘power’ is used to denote the rate of energy transfer, and so the watt refers to a unit of power and a rate at which heat is transferred. However, both the joule and the watt are quite small amounts, and it is usually more practical to encounter units of heat and heat transfer in terms of kilojoules (kJ) and kilowatts (kW) respectively.
Transfer of heat As previously mentioned, there are three methods by which heat can be transferred from one point to another: • Conduction – the transfer of heat between bodies in contact (external
conduction) or from one part of a body to another part (internal conduction). Good conductors transfer heat more readily than poor conductors.
• Convection – the transfer of heat to or from a fluid (gas or liquid) moving over a solid surface. This is heat that is conveyed from place to place by the actual motion of the heated fluid.
• Radiation – the transfer of heat across space by electromagnetic rays identical to light rays in nature, but of a lower frequency. A hot body radiates heat in all directions in straight lines, and in consequence the radiation may be cut off by a screen (as in a cloud obscuring the sun). The rays do not appreciably warm the medium through which they pass, that is, the rays of the sun do not heat the air but, in fact, heat small particles of dust and moisture in the atmosphere, which then pass on the heat to the air by convection.
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Hot Water Heating
THERMAL ENVIRONMENT In simplistic terms, if it is cold outside then, to maintain an acceptable environment within a space, we must put sufficient heat into the space to maintain a comfortable temperature. The amount of heat we add will depend on: • The internal temperature we wish to maintain • The external temperature • The type of building fabric and its area which will determine the rate at
which we lose heat from the space • The volume of the space and the number of times that volume of air will be
changed by natural means by leaking out from gaps around windows and doors etc.
In principle, all these are simple to determine, but in practice, are difficult to obtain. The whole subject of heat losses from buildings is complex and has been addressed many times by many learned bodies. This is such a complex area that we will not attempt to delve into a great deal of detail in this document. But we will outline some of the main points and then return to a simple approach to outline the basics. Let’s begin with the human body (Fig. 2).
33oCM – Metabolic Heat α Food Energy Activity = E ± C ± R E – Loss by Evaporation in Perspiration, affected by:- Air Movement. Air Humidity. C – Convection Loss from skin surface, affected by:- Air Temperature. Air Movement. C1 – Conduction Loss (Minimal) by:- Direct Contact. R – Loss by Radiation affected by:- Radiant Temperature. Air Temperature. Differential Surface Temperature of Adjacent bodies.
R – Radiation (50%) 30oC
21oC
E – Evaporation (20%)
C – Convection (30%)
Optimum Skin Temperatures
Core36oC37oC
18oC
Fig. 2 Body Heat Balance
C1 – Conduction
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Hot Water Heating
From this we can see that the heat balance of the human body to its surroundings is affected by: • Air temperature • Radiant temperature air movement • Air humidity All these factors will influence the degree of ‘comfort’ experienced within the space. It has been customary to base heat loss calculations on the difference between design external air temperature and design internal air temperature. In mainland Europe, the dry resultant temperature is often used, which is based on a globe temperature with a diameter of 150 mm; whilst in the UK in recent years, yet another scale is sometimes used – Environmental Temperature (Tei). This is claimed to produce a better index of comfort and is dependent upon the dry bulb internal air temperature (Tai) and mean surface temperature (Tm). Equation 1 can be used to give a close approximation of environmental temperature.
mei ai2 1T = T + T Equation 13 3
Where: Tei = Environmental temperature Tm = Mean surface temperature Tai = Dry bulb temperature With this scale, the type of system used influences the heat loss calculations, that is, different formulae exist for calculating heat losses if the system is a convective heating system or a radiant heating system.
Heat losses As was stated earlier, in order to outline the basic concepts, we will discuss the principles of heat loss calculations in simplistic terms by considering dry bulb temperatures. Whatever the type of building or heating system, the first step of the design is the estimation of heat losses. Fig. 3 shows the major factors involved and you will note that there may be some heat gains to be taken into account. In northern European latitudes these can generally be ignored (except if some form of supplementary heating exists, or there is some exceptional machinery) when designing a heating system. However, if an air conditioning system is being proposed, the effects of heat gains can be significant and must be taken into account.
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Hot Water Heating
1 1
1
1
B A
C
D
1
2
1
Heat Loss 1. Fabric losses through roof, walls, floor and windows 2. Air Changes – Ventilation
Heat Gains A. Solar Radiation B. Personnel – dependent on number/activity C. Lighting D. Heating/Machinery
Fig. 3 Heat Losses and Gains The fire (item D in Fig. 3) can be used to explain the terms ‘background’ heating and ‘full’ heating. If supplementary heating is provided by a fire for example, the system would be designed to provide background heating, that is, a lower internal design temperature supplemented by a fire. If no fire (or suchlike) exists, a full heating system would be provided. The mechanics of carrying out heat loss calculations involve taking each room or space in turn and estimating the amount of heat necessary to maintain a steady internal temperature against a steady external temperature. These calculations are based on ‘steady state’ conditions, that is, the running load once the building is up to temperature; no allowance is made at this stage for actually getting the space up to the desired internal temperature in the first place. Depending on the use of the building, an allowance will be made at the end of the calculation to cater for either ‘continuous’ or ‘intermittent’ operation of the heating system.
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Hot Water Heating
To take extreme cases, a church hall may only be used once or twice a week. So the heating system must be capable of bringing up the space temperature in a reasonable time, as well as maintaining steady conditions. Conversely, a hospital ward has continuous use and the system will be operating at all times during seasonal months. Initial heat up can be carried out over a comparably long period and the system designed simply to maintain steady conditions.
The heat loss calculation The heat loss calculation is in two parts, fabric loss and ventilation loss.
Fabric loss The heat loss due to the conduction of heat through the materials of construction of the walls, floor, ceiling, windows etc. next to unheated spaces (for example, an outside wall); this being termed ‘fabric’ or ‘transmission’ loss. Heat transfer coefficients (‘U’ values) are recorded and available for each type of construction material. These denote the rate of heat flowing through unit area of the material in question, in unit time, for unit temperature difference. The derived unit for transmission loss is watts per square metre per degree Celsius (W/m2oC). Thus the area (m2) of each separate type of material (glass, brick, cavity wall etc) must be calculated and multiplied by the appropriate U value. The selection of the U value may be influenced by the aspect of the building, that is, whether it can be considered ‘normal’, ‘sheltered’ or ‘exposed’.
Ventilation loss The second part of the calculation involves the assessment of the amount of heat necessary to warm the volume of air passing through the space being heated, again to maintain a steady internal temperature against a steady external temperature. This is known as ‘natural air change’ or ‘infiltration’. Gaps exist around windows and doors and, depending on the use of the room in question, doors are opened and closed at regular intervals. All these factors mean that the volume of air in the space is changed at a certain rate every hour. Research into the average rate of air change for a room has resulted in empirical figures for the number of air changes per hour being available for designers (see Table 1). It must be stressed that natural air change is due to the inadequacies of building construction methods and happens ‘naturally’ and is not, strictly speaking, controlled.
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Hot Water Heating
Table 1 Typical Empirical Air Filtration Rates for Normal Winter Heating excluding Mechanical Ventilation
Building Air changes per hour
Art gallery 1
Bank 1 – 1½
Church ¼ – ½
Restaurant ½
Flats
- Living rooms 1
- Bedrooms ½
Offices 1
Shops ½ – 1
To further illustrate this, you will see from Table 1 that the average number of air changes per hour for a UK living room is 1. In Australia, where, due to the climate, building construction can be ‘looser’, it is more usual to allow 4 changes per hour. Conversely in Canada, again due to climatic needs, construction is ‘tighter’ and only ½ an air change per hour is normal. Where mechanical ventilation is employed, then the known rating of the fans must be used to determine the heat losses due to air changes within the space. The specific rate of heat loss due to natural air change is obtained from Equation 2.
o
3
3
o
. .p
.
.
p (at constant pressure)
q V c Equation 2
Where :
q Specific rate of heat loss due to air change kW / C differential temperature
V Rate of ventilation m / s
Density of air kg /m
c Specific heat of air kJ/kg
= ρ
=
=
ρ =
= C
We can use standard values for air density and specific heat of 1 kg/m3 and 1.2 kJ/m3oC respectively. By using these standard values, we can simplify Equation 2 into Equation 3.
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Hot Water Heating
.q 0.33 N V Equation 3
o
3
3
.
Where :
q Specific rate of heat loss due to air change W / C of differential
N Number of air changes per hour m /h
V Room volume m
=
=
=
=
Design temperatures The heat loss calculations must be based on maintaining a pre-determined design internal temperature against a design external temperature. The design internal temperature will depend upon the geographic location, type, use, and occupancy of the building. The design internal temperature depends on the average external winter conditions applicable to the specific geographic location of the building. Local meteorological records and heating textbooks will provide information. The difference between internal and external temperatures is the ‘design temperature difference’ (ΔTdes).
Example 1 Calculate the heat loss from a typical building Let us consider the simple basic heat loss calculations for the building shown in Fig. 4.
50m 150m
1.50m
4m Glass*
Door 3m
80m
2m * Same amount of glass on opposite side Fig. 4 Building to be heated
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Hot Water Heating
1. Design data: Internal design temperature = 21oC External design temperature = -5 °C Number of air changes (N) = 1.5 U values (W/m2oC) Walls 0.65 Floor 0.11 Roof 0.35 Windows (double-glazed) 1.80 Door 2.00 Note: Above values are typical values only. Actual ‘U’ values for the application must be obtained prior to any calculations. 2. Specific fabric heat loss Let us first calculate the fabric areas (m2): Windows 80 x 1.5 x 2 = 240 m2
Door 3 x 2 = 6 m2
Roof 150 x 50 = 7 500 m2
Floor 150 x 50 = 7 500 m2
Walls (150 x 4 x 2) + (50 x 4 x 2) - (Door area + window area) (150 x 4 x 2) + (50 x 4 x 2) - (6 + 240) 1 200 + 400 – 246 = 1 354 m2
We can now calculate the total fabric specific heat loss from the respective fabric areas and thermal transmission losses (U Values). Area (m2) x U value = W/ oC Walls 1 354 x 0.65 = 880 Floor 7 500 x 0.11 = 825 Roof 7 500 x 0.35 = 2 625 Windows 240 x 1.80 = 432 Door 6 x 2.00 = 12
Total specific fabric heat loss (∑AU) = 4 774 W/ oC
1. Specific air infiltration loss ( ) .q
Volume of building (V) = 150 x 50 x 4 = 30 000 m3
Air changes (N) = 1.5 per hour
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Hot Water Heating
The specific rate of heat loss due to air change is found from Equation 3:
o
.
.
.
q 0.33 N V Equation 3
q 0.33 x 1.5 x 30 000
q 14 850 W / C of differential
=
=
=
4. Design temperature difference (ΔTdes) Internal design temperature = 21 oC External design temperature = - 5 oC Design temperature difference = 26 °C 5. Total heat loss from building We can now determine the building total heat loss by adding the above information to Equation 4.
Total heat loss = (∑AU+.q ) x ΔTdes Equation 4
Where: Specific fabric loss (∑AU) = W/°C of differential
Specific air infiltration loss .q = W/°C of differential
ΔTdes = °C differential
Total heat loss (Q ) = (4 774 W/.
TOT oC + 14 850 W/ oC) x 26°C
= 510 224 W By dividing the total heat loss with the volume of the building we can arrive at a coefficient for the heat loss per unit volume:
Heat loss per unit volume = 510 224 W 30 000 m3
= 17 W/m3
Nominal values for buildings of different sizes can be used to give a general indication of heat losses and therefore probable heating requirements. These are as given below: Dwellings 40 to 60 W/m3
Buildings up to 3 000 m3 30 to 40 W/m3
Buildings above 3 000 m3 15 to 30 W/m3
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Hot Water Heating
It must be stressed that the above figures are only nominal values intended to give a rough estimation of heating loads; there is no substitute for carrying out proper heat loss calculations.
Example 2 Environmental temperature method of calculation We mentioned earlier that another scale for comfort is sometimes used in the UK, that is, environmental temperature (Tei). It is interesting to see how the Tei heat loss calculations compare with those in Example 1, and on the assumption that a convective heating system is to be installed. 1. Fabric specific heat loss Area (m2) x U value = W/ oC Walls 1 354 x 0.65 = 880 Floor 7 500 x 0.11 = 825 Roof 7 500 x 0.35 = 2 625 Windows 240 x 1.80 = 432 Door 6 x 2.00 = 12
Total specific fabric heat loss (∑AU) = 4 774 W/ oC 2. Design temperature difference (ΔTdes) Internal design temperature = 21 oC External design temperature = - 5 oC Design temperature difference = 26 °C 3. Fabric heat loss Fabric heat loss (∑QF) = ∑AU x ΔTdes
∑QF = 4 774 W/ oC x 26 °C ∑QF = 124 124 W
4. Applying a factor for environmental heat loss The environmental heat loss factor is based on the ratio of the total fabric heat loss (∑QF) to the total area comprising the building fabric ∑A, in units of watts per square metre. ∑A = 1 354 + 7 500 + 7 500 + 240 + 6 = 16 600 m2
ΣQF = 124 124 WΣA 16 600 m2
ΣQF = 7.5 W/m2
ΣA
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Hot Water Heating
From the environmental heat loss factor ( FQ A∑ ∑ ), a temperature correction factor (TCF) is obtained from Table 2: Table 2 The Temperature Correction for Environmental Heat Loss
FQ A∑ ∑ (W/m2)
TCF (oC)
FQ A∑ ∑ (W/m2)
TCF
(oC) FQ A∑ ∑ (W/m2)
TCF (oC)
5 1.0 17 3.5 29 6.0
6 1.3 18 3.7 30 6.3
7 1.5 19 4.0 31 6.5
8 1.7 20 4.2 32 6.7
9 1.9 21 4.4 33 6.9
10 2.1 22 4.6 34 7.1
11 2.3 23 4.8 35 7.3
12 2.5 24 5.0 36 7.5
13 2.7 25 5.2 37 7.7
14 2.9 26 5.4 38 7.9
15 3.1 272 5.6 39 8.1
16 3.3 28 5.8 40 8.3
From Table 2, For an environmental heat loss factor of 7.5 W/m2, the TCF is 1.6oC. 5. Calculating the air infiltration differential temperature (ΔTvent) The TCF is now added to the (ΔTdes = 26oC) to give a new air infiltration differential temperature (ΔTvent).
ΔTvent = ΔTdes + TCF ΔTvent = 26oC + 1.6oC ΔTvent = 27.6oC
6. Calculating the air infiltration loss
Specific air infiltration loss ( ) .q
Volume of building (V) = 150 x 50 x 4 = 30 000 m3 Air changes (N) = 1.5 per hour From Equation 3:
o
.
.
.
q 0.33 N V Equation 3
q 0.33 x 1.5 x 30 000
q 14 850 W / C of differential
=
=
=
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Hot Water Heating
Using the new air infiltration differential temperature (ΔTvent), the total air
infiltration loss ( ) is calculated from this and the specific air infiltration loss
( ) as shown in Equation 5.
.Q
.q
o
o
. .vent
.
.
vent
Total air infiltration loss (W)
Specific air infiltration loss (W/ C)
=Air infiltration differential temperature ( C)
Q q T Equation 5= Δ
Where :
Q
q
T
=
=
Δ
Q = 14 850 W/oC x 27.6oC Q = 409 860 W
7. Calculating the total heat loss
The total heat loss = ∑QF + Q .
= 124 124 W + 409 860 W = 533 984 W 8. The nominal value for the heat loss for this building By dividing the total heat loss with the volume of the building we can arrive at a coefficient for the heat loss per unit volume:
Heat loss per unit volume = 533 984 W 30 000 m3
= 17.8 W/m3 This value for Example 1 was 17 W/m3. It can be seen that, by using the environmental temperature method of calculation, the nominal heat loss has increased by 4.7%.
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Hot Water Heating
BASIC ELEMENTS OF A HOT WATER HEATING SYSTEM It would now be as well to look at some of the basic items present in a hot water heating system, and also to clarify some basic definitions. Conventional hot water heating systems are categorised as shown in Table 3: Table 3 Categories of Hot Water Heating Systems
Category Temperature Range (oC) Typical Flow/Return Temperature & ΔT (oC)
LTHW Low temperature hot water
Up to 100
80/70 : ΔT 10 82/71 : ΔT 11 85/65 : ΔT 20 80/60 : ΔT 20
MTHW Medium temperature hot water
100 - 120 *120/90 : ΔT 30
HTHW High temperature hot water
Above 120 *150/110 ΔT 40
* Other temperatures and ΔT’s are used – tends to be system specific No mention of pressure is made when categorising a heating system. LTHW systems nowadays generally operate with a pressurisation unit. Older systems still exist which operate under a static head supplied by an open top feed and expansion tank. MTHW and HTHW systems require pressurisation to stop the water from boiling at the higher temperatures used.
Feed and expansion As a general guide, the expansion of water from 10oC to 100oC is 1/23 of its original volume, and the system design must be able to cater for this to prevent excess pressure and, ultimately, damage to the system. Table 4 shows expansion percentage of water from 4oC;
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Hot Water Heating
Table 4 Expansion Percentage versus Water Temperature oC oF Expansion
% (e) oC oF Expansion
% (e)
4 39 0.00 60 140 1.71
8 46 0.02 62 144 1.82
10 50 0.03 66 151 2.05
12 54 0.06 68 154 2.15
16 61 0.11 70 158 2.26
20 68 0.18 74 165 2.52
22 72 0.23 78 172 2.76
24 75 0.28 80 176 2.91
28 82 0.33 82.2 180 3.07
30 86 0.44 86 187 3.30
32 90 0.50 88 190 3.44
36 97 0.64 90 194 3.60
38 100 0.70 94 201 3.87
40 104 0.78 96 205 4.04
42 108 0.86 98 208 4.19
46 115 1.03 100 212 4.35
50 122 1.21 105 221 4.77
52 126 1.30 110 230 5.15
56 133 1.50 115 239 5.62
58 136 1.60 120 248 6.01
Fig. 5 shows a simple heating system. Historically, a feed and expansion tank would be fitted to allow for expansion. However, all modern systems now use a pressurisation unit of one form or another to both provide the initial filling and pressurisation of the system, and then to cater for the expansion thereafter.
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Hot Water Heating
Older Systems
Open Vent
Overflow
Fall Pump
Cold Feed
Heat Emitter (Radiator) Pressurisation Unit Boiler
Fig. 5 Simple Heating System A typical pressurisation unit (courtesy of Grundfos), and its components, is shown in Fig. 6.
Primary features: - Pump (can be duty/standby) - Break tank - Pressure switch - Expansion vessel - Non return valve - Pressure gauge
Fig. 6 Typical Pressurisation Unit The benefits of fitting a pressurisation unit are many. Far less space is required and a traditional feed and expansion tank necessitates a location at the highest part of the building, which will also need a supporting structure – all of which can be difficult and costly. Other major benefits of using a pressurisation unit include reduced maintenance and less corrosion in the system as, with a sealed system, once treated there is no way for additional oxygen to enter the system.
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Hot Water Heating
Fig. 7 shows the principle of operation of the expansion vessel;
Diaphragm A C B
Water System
Air or
Nitrogen Cushion
A. When system is filled, no water enters tank when cushion & water pressure are in
equilibrium. B. As temperature increases, diaphragm moves to accept expanded water. C. When water rises to maximum, full acceptance of expansion is achieved. Fig. 7 Closed Diaphragm Expansion Vessel The sizing of the expansion vessel is generally done by the supplier, however, the information required is generally as follows;
• System working pressure • The highest point within the system above the pressure vessel (B) • Cold fill pressure (normally B + 0.35 bar) • The maximum allowable system working pressure • The volume of water within the system (rough estimate if not known is
12 litres per kW) • Glycol content if a chilled water system • On high resistance systems, the pump discharge pressure may be
required
Air removal It is vital that air in a water system is removed since it will cause corrosion and retard the circulation of the water. Manually operated air vents may be useful during initial filling of the system, but these will not remove any air or other gases which may collect during normal running. For these reasons, it is preferable to use an automatic air vent at strategic points in the piping system. Fig. 8 shows an AE30 Automatic Air Vent for water systems;
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Hot Water Heating
Fig. 8 AE30 Automatic Air Vent On medium to large systems the tendency is to fit a proprietary air and dirt separator, an example of which is shown in Fig. 9;
Fig. 9 Coalescing Dirt & Air Separator Coalescing style air and dirt separators work by slowing the movement of system fluid through a large cross-section of the air separator tank. The coalescing tubes allow the slow-moving micro bubbles to cling to the stainless steel tubes and coalesce, or join together. The bubbles then rise to the top of the air separator, where they are vented through an automatic air vent. In addition, dirt particles are directed down through the non-turbulent zone and stored in the dirt chamber at the bottom of the unit, where they can be removed periodically.
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Standard Symbols and Abbreviations
Abbreviations Heating Surfaces/Emitters AAV Automatic air vent ConvectorsF Flow Natural Fan Assisted FA From above FB From below F&E Feed and expansion Embedded PanelsHL High level Floor Ceiling HTHW High temperature hot
water LL Low level LSV Lockshield valve Pipe CoilsLTHW Low temperature hot High Level Low Level
water MTHW Medium temperature
Hot water Radiant Strip Radiator NRV Non return valve OV Open vent R Return Radiant PanelsTA To above Industrial Single Industrial Double TB To below WV Wheel valve Wall Mounted Ceiling Mounted
Piping symbols Unit Heaters Horizontal Downward
Pipes in roof or aboveceiling
Pipes at high level
Vertical drop or rise
Pipes at low level
Pipes below floor
Direction of flow
Gradient (rise or fall should be indicated)
Isolating Valve
Check Valve
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Hot Water Heating
HEAT EMITTERS In our simple building heating example (Example 1), we have determined how much heat we need to put into the building to maintain comfortable internal temperatures. In order to do this, it is necessary to install some kind of heat emitting equipment, and it is therefore appropriate to discuss the various types available for such purposes. Under the overall context of conventional heating equipment, the range of heat emitters can be divided into three natural groups: a) Natural convective b) Forced convective c) Radiant
Natural convective
Radiators Despite their name, the heat emission from radiators is very largely achieved by convection. Various types and materials are available as well as differing heights, lengths and depths. Their heat output is again given in terms of watts per square metre, and the surface area of each specific radiator will be given in the manufacturer’s data.
Natural convectors Generally comprise finned or grilled tubing within a cabinet which has a grilled outlet at the top. Outputs can vary from 200 W to 20 kW. Finned tubing can be supplied in continuous from, when the output is stated in watts per metre run.
Forced convective
Fan convectors Similar to natural convectors except that a fan within the casing forces air through the unit resulting in a high heat output per volume of space occupied by the unit. Outputs can vary from 1.5 kW to 25 kW.
Unit heaters A unit fitted with a heat exchanger and a large fan to give high air volumes and wide throws. Outputs from 3 to 300 kW.
Radiant
Radiant panels Consists of tube attached to a radiating surface. Very useful for ‘spot heating’ local areas such as personnel in a large warehouse. Outputs are again given in terms of watts per square metre.
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Hot Water Heating
Radiant strip Long runs of pipe attached to a radiant surface, with heat outputs quoted in terms of watts per metre run. Two other types of heat emitting systems are worth mentioning:
Under-floor heating Hot water coils are embedded in floor screed, the floor surface being heated to about 23 to 26°C. Heat is emitted largely in radiant from, and the maximum water flow temperature must generally be kept below 60°C. Higher temperatures can cause excessive expansion and result in cracking of the floor screed.
Heated ceilings This type of system originated with small bore pipes embedded in the structural slab and then plastered over. These are fed by hot water at relatively low temperature (37 to 48°C) to prevent excessive expansion and cracking of the plaster.
A later version had the piping encased in a sleeve enabling higher temperature (71 to 82°C) water to be used.
The latest types make use of light metal trays perforated for acoustical effect and insulated above, being either clipped to a grid of heating pipes or independently suspended below a series of pipe coils.
Heated ceiling systems emit heat in radiant form.
Emission figures
A general indication has been given on the sort of output that can be expected from various types of heat emitter. We need to consider this a little further, taking the radiator as our example.
The actual heat emission from radiator depends not just on the emissivity figures in W/m2 given by the manufacturer, but also on the temperature difference between the mean water temperature of the radiators and the desired room temperature. Some radiator manufacturers will quote emission data on a specified temperature difference, perhaps 50°C. Any variation in the design temperature difference means that a correction factor must be applied to give the correct heat output.
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Hot Water Heating
PIPING SYSTEMS If we are faced with the task of designing a heating system, we will have calculated the heat losses, and selected and positioned our heat emitters. We now require a pipework system to connect all the emitters to our heat source such that they receive adequate amounts of flow water at the required temperature, transfer the heat to the space being heated, and return the lower temperature water exiting from the heat emitter back to the heat source, commonly called a boiler. Types of piping systems for LTHW heating installations can be categorised into one pipe, two pipe direct return, two pipe reverse return, or combinations and variations of these.
Gravity systems With the advent of reliable, relatively cheap, small pumps, it is true to say that very few, if any, gravity systems in their true sense are designed today, so we will not cover them here.
Pumped systems In a pumped (or forced) system, the water is mechanically forced around the circuit. In nearly all cases the force is generated by a centrifugal circulating pump driven by an electric motor. Consequently, compared with a gravity system, we can circulate the same quantity of water through much smaller diameter pipes. With smaller pipes, there is less water in the system which allows a more rapid heat up and makes the system much more responsive to control. The higher the circulating force developed by the pump, the smaller need be the pipes to deliver a given quantity of water. From a cost viewpoint, the smaller the circulating pipes the cheaper the installation. There is, however, a limit in terms of just how small a diameter these pipes can be. As the pipe bore decreases to carry a given quantity of water, the higher the velocity needs to be through the pipe. This can cause unacceptable noise and, in extreme cases, corrosion in the pipes and fittings. Water velocity is an important criterion that is used when carrying out pipe sizing, as we shall see later.
Types of circuit If we have established our heat losses, selected our heat emitters and a boiler and pump, we now need to concentrate on sizing the pipe to carry the circulating water from the boiler to the radiators. We also need to return water from the radiators back to the boiler. This is called the piping system or circuit or sometimes ‘network’ on a very large system. There are several types of circuit which can be employed.
One pipe systems The simplest system is known as a single pipe series loop, as depicted in Fig.10.
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Hot Water Heating
Here, water is circulated through one pipe to each radiator in turn. Effectively, the radiators form part of the circulating system. Closing off one radiator would shut down the complete system, and this circuit is rarely used in practice. But it does lead us to the next type of circuit, which is the true one pipe system, as depicted in Fig.11.
Fig. 10 Series Loop System
Fig. 11 One Pipe System The one pipe system is a variation on the series loop system in that the radiators are now connected to the single pipe by short risers via diverting tees in the main circulating line. With a one-pipe system, cooled water emerging from each radiator mixes with the warmer water which bypasses the radiator. The water therefore becomes progressively cooler after each radiator is served. To compensate for this, progressively larger radiators are installed along the system in order to give the same output, as depicted by Fig. 12.
Pump Emitters/Radiators
Boiler Pressurisation Unit
Diverting Tee
A B
Pump
Pressurisation Unit Boiler C D
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Fig. 12 Radiator Water Temperatures on a One-Pipe System In this example, radiator B now accepts water at a lower temperature than radiator A. This essentially means that radiator B will have a lower mean temperature than radiator A, and if it were the same size as radiator A would emit less heat because of it. Consequently, radiator B will be slightly larger than radiator A to compensate for this.
Two pipe system – Direct run On such a system, as shown in Fig.13, water which serves a radiator returns direct to the boiler and does not mix with, and thus cool, the flow water passing to the next radiator in the circuit. Each radiator closes a circuit between the flow and return pipes, and as a result, water will be supplied to each radiator at about the same temperature. Water will also return from each radiator at about the same temperature (assuming the same conditions). This means that the mean water temperature of the various radiators is more uniform than those in a one-pipe system, enabling radiators of the same size to be used for identical outputs. Also, pipes can be reduced in size along the circuit, as discussed below in relation to Fig.13.
Fig. 13 Two Pipe Direct Return The system at point X must carry sufficient water to serve radiators A, B and C; at point Y, only enough water to serve radiator C is needed. This type of system does have one disadvantage. Since each radiator forms a circuit with the flow and return pipes, the lengths of each radiator circuit will
Radiator Mean Water
Temperature 75oC
Radiator Mean Water
Temperature
A B 80oC 78oC
73oC 70oC 68oC
80oC 78oC 75oC
A CB
X Y
Pressurisation Unit
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vary greatly according to their distance from the boiler. This can be seen by referring to Fig. 14.
Fig. 14 Radiator Water Temperatures on a Two-Pipe System It can be seen from Fig. 14 that circuits ABCDE and AFGHE are of different lengths, and this causes a tendency for the water to pass through that circuit of least resistance. This will normally be the shorter circuit, that is, the radiator closest to the boiler and pump, in this example circuit ABCDE. To prevent this, balancing valves must be fitted to radiators and adjusted; those nearer the boiler being more closed than those further away, which may be wide open. We will discuss the finer arts of balancing in more detail later, for this is a very important part of commissioning a successful system.
Two pipe system – Reverse return The short-circuiting problems associated with the two pipe direct return circuit can be dealt with by the reverse return two-pipe system, which overcomes the need for balancing. Each radiator circuit is of identical length, that is, the radiator with the shortest flow from the boiler has the longest return and vice versa, thus ensuring equal resistances and prevent short-circuiting and the need for balancing valves (within the limitations of commercial pipe sizes). This can be seen in Fig. 15.
Fig. 15 Two Pipe Reverse Unit
F B 80oC 80oC
31
Radiator Mean Water
Temperature 75oC
70oC 70oC
80oC
Radiator Mean Water
Temperature 75oC
Flow 80oC
A
70oC D H
E 70oC Return
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The disadvantage of the reverse return system is that additional pipework, fittings and installation costs are required above that of the direct return system. In a square building, direct and reverse return systems would probably have similar piping costs, but the reverse system would not need balancing. In a long narrow building, the cost of a direct return system would probably be quite a lot less than a reverse return system, even allowing for balancing. Obviously, this is just a generalisation, each building would need consideration based on its own merit. In practice, the reverse return system is rarely used, with the exception of plantrooms where it is used to balance the flow of water through heat exchangers, thus, ensuring equal load on each unit.
Flows of liquids through pipes More information on this subject is given in the Learning Centre, Modules 4.1, and 10.2.
The effect of the fluid on flow Liquids fall into one of two categories: • Newtonian fluids • Non-Newtonian fluids Newtonian fluids obey Newton’s law of viscosity, which states that the shear stress developed by a fluid in motion equals and is linear to the product of the fluid’s dynamic viscosity and its velocity gradient from the pipe wall. Non-Newtonian fluids are much more viscous, such as plastics and colloidal solutions, clays, milk, cement and slurries etc. Water is classified as a Newtonian fluid and its flowrate and velocity are directly proportional (with constant density), as depicted in Equation 6. It is also incompressible and its behaviour, when flowing through pipes, is generally predictable.
3
2
.
.
m u A Equation
Where :
m Mass flowrate (kg/s)
= Density (kg/m )
u = Velocity (m/s)
A Pipe area (m )
= ρ
=
ρ
=
6
Factors which affect the flow of liquid in a pipe include: Average velocity of the liquid (u)
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Dynamic (or absolute) viscosity of the liquid (μ) Density of the liquid (ρ) Diameter of the pipe (D) All of these factors can be brought together in one dimensionless quantity to express the characteristic of any one fluid passing through a pipe. This is known as the Reynolds’s number (Re), and is quantified by Equation 7. The Reynolds’s number indicates whether fluid flow is laminar or turbulent or at some indeterminate point in between. It is generally accepted that Reynolds’s numbers below 2000 indicate laminar flow, between 2000 and 4000 indicate a transitional state, and above 4000 indicate turbulent flow. Each of these different flow types has a different affect on how friction is created in the pipe, and the Reynolds’s number can be used to determine the pipe friction factor, knowing the degree of smoothness of the pipe wall.
3
e u D
R E
Where :
Re Reynold's number (dimensionless)
= Density (kg/m )
u = Velocity (m/s)
D Pipe diameter (m)
Dynamic viscosity (kg/m s)
ρ=
μquation 7
=
ρ
=
μ =
We can see from Equation 7 that, for all other things being equal, the greater the fluid’s dynamic viscosity (μ), the lower the Reynolds’s number. The greater the dynamic viscosity, the more the fluid drags on the pipe walls, and laminar flow tends to be the result, as seen in Fig.16. For less viscous fluids like water, less drag occurs at the pipe wall, the flow breaks up and turbulent flow tends to result, with a much flatter velocity profile, as seen in Fig.17.
Fig. 16 Laminar Flow Fig. 17 Turbulent Flow The ‘drag’ at the pipe wall is the result of friction between the moving fluid and the pipe wall. For any one pipe, we can imagine that a more viscous fluid, perhaps syrup, would incur more friction than water.
Flow Parabolic Velocity Profile
Flow Velocity Profile
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Nevertheless, even though water has a relatively low viscosity, it still creates friction as it passes the pipe wall; this friction removes some of the energy that was supplied to the water to move it through the pipe. Pressure energy is lost as the water travels along the pipe, and the system loses ‘head’ or pump pressure as a result. It isn’t just dynamic viscosity that affects the Reynold’s number. The density of the fluid is also considered in the Reynold’s number equation. If we compare air to water, we can see that water might have a dynamic viscosity of about 1 kg/m s, while the dynamic viscosity of air at the same conditions might be about 0.018 kg/m s, a factor of some fifty-five times less. This might cause us to think (by looking at Equation 7) that the Reynold’s number for air is a lot less than that for water. But this is not actually the case if we compare both fluids travelling at the same speed down the same sized pipe. If we now consider the density of both fluids, we can see that the overall result is somewhat different. The density of water at the same conditions would be about 1 000 kg/m3, whereas the density for air would be about 1.2 kg/m3, a factor of eight hundred and thirty times less. This would mean that, for the same velocity and same size pipe, the Reynold’s number for water would actually be 830/55 = 15 times larger than that for air. The relationship between dynamic viscosity and density is termed kinematic viscosity, and is shown by Equation 8.
fluid dynamic viscosity ( )kinematic viscosity (v) = Equation 8
fluid density ( )μ
ρ
By taking account of Equation 8, we can derive another Equation from Equation 7, which considers the Reynold’s number in terms of the kinematic viscosity, as shown by Equation 9.
2 3
u DRe = Equation 9
vWhere:
Re = Reynold's number (dimensionless)
u = Fluid velocity (m/ s)
D = Pipe diameter (m)
v = Kinematic viscosity (m /s )
Using Equation 9, we can see how the Reynold’s number for water and air differ for the same velocity of 3 m/s and same size DN50 pipe. For WATER, From Equation 8
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1kinematic viscosity (v) =
1 0000.001=
From Equation 9,
3 x 0.05
Re = 1500.001
=
For AIR, From equation 8
0.018kinematic viscosity (v) =
1.20.015=
From Equation 9,
3 x 0.05
Re = 100.015
=
We can see from these results that, for the same sized pipe and the same velocity, the Reynold’s number for water would be fifteen times larger. However, in practice, recommended water velocities would be, on average, about 2 m/s, whereas a recommended maximum air velocity is about 9 m/s.
The effect of the pipe on flow The pipe itself also has an affect on the fluid passing through it. A smooth bored pipe will allow fluid to flow more easily than a rough bored pipe, and these different frictional effects should be considered during the design process. We do this by referring to a ‘friction factor’ (or coefficient of friction) which accounts for the relationship between the roughness of the considered pipe and the Reynold’s number for the fluid. In straight pipes, the pressure loss due to friction can be determined by use of Equation 10, sometimes referred to as the D’Arcy equation:
2
2
f
f
4 f L uh E
2 g D= quation 10
Where :
h Head loss due to friction (m)
f = Coefficent of friction (Dimensionless)
L = Length of pipe (m)
u = Fluid velocity (m/s)
g = Gravitational constant (9.81 m/s )
D = Pipe diamater (
=
m)The above relates to straight pipe but, inevitably, pipe fittings like valves,
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36
reducers, bends, tees, elbows etc will be part of the pipework, and the effect of these also need to be considered. A fitting will invariably produce higher pressure drops than the pipe it replaces because they invariable cause the fluid to change direction or travel faster, which increases the resistance to flow. It is possible to take into account the higher friction losses of such fittings by considering them in terms of having an ‘equivalent length’ of straight pipe, which can be added to the actual pipe length. These equivalent lengths (le) of pipe are multiplied by a velocity head factor (k) dependent upon the type of fitting or piece of equipment in question – we shall see how this works later on in the text. Liquid flow in pipes is a very complex subject indeed, and we have done little more than gloss over the surface. Fortunately, we don’t need to begin our pipe sizing calculations from first principles, since tables are available that inform us of friction loss, equivalent pipe length and associated velocity head factors, and velocity. Table 5 shows the relationship between pressure drop, pipe diameter, mass flowrate, equivalent pipe lengths, and velocity. Table 6 gives typical velocity head factors (or friction factors = ‘k’ values) for different types of fittings and equipment used in the industry. Different tables exist for different grades of pipes and fluid temperatures.
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Table 5 Flow of Water at 75oC in Steel Pipes
Note: dp/l = Pressure loss per unit length (Pa/m) m = Mass flow rate (kg/s) le = Equivalent length of pipe in metres for k = 1.0 v = Water velocity (m/s) k = Velocity pressure factor, see Table 6 Different tables exist for the various grades of steel tubes and for copper tubes. A multiplier is also available for higher and lower water temperatures. e.g. 150oC and 10oC.
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Table 6 Values of Velocity Head Factor ‘k’ for Pipe Fittings & Equipment
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39
PIPE SIZING
Introduction Pipe sizing consists essentially of selecting the correct pipe diameter for the amount of water that it is to carry, such that it is economically sized to prevent an excessive pressure drop which would result in high pumping costs and high velocities resulting in noise. Nor should the pipe be larger than is necessary in order to keep capital outlay to an optimum. In practice, a number of assumptions are made to enable us to carry out preliminary pipe sizing. These include: -
1. Adoption of a pressure drop within the range of 200 to 450 Pa/m (Lower values for small systems).
2. A check to see that we do not exceed water velocities of 1 to 3 m/s and perhaps up to 4 m/s for large system distribution mains. (Note - From Table 6, it will be seen that these flow velocities will not be exceeded if we adopt the 200 to 450 Pa/m pressure drop as a basis for preliminary pipe sizing.
3. An allowance for heat emission from the pipes serving the heat emitters. This may be between 5% and 25% of the unit's heat output depending on the length of pipe, whether it is insulated etc.
4. An allowance for fittings etc. This is usually an additional 10% - 30% on the actual pipe length, depending on the ratio of fittings to pipe length.
Using Tables, such as Table 5, we can then carry out preliminary pipe sizing. That may be sufficient to prepare initial cost estimates and for small, simple systems. However, it is usual to use the information based on the assumptions in the preliminary pipe sizing exercise, to carry out final pipe sizing including proportioning the heat emission of the mains between the various branches.
Step-By-Step Pipe Sizing
Preliminary pipe sizing
Example 3 Fig.18 shows a fairly straightforward LTHW system in diagrammatic form which we will use to demonstrate the various steps in the pipe sizing routine. We must assume that we have carried out the heat loss calculations to provide internal comfort conditions against average outside winter temperature. We intend to use radiators as the heat emitters and, based on the heat loss figures for each room in the building, these have been selected so far as the amount of heat surface is concerned, to provide sufficient output to maintain the required room temperature for the room in question at the design outside ambient temperature. The emission (in kW) is marked on each radiator. We have also determined that our LTHW system will be designed to provide flow temperature water at 80oC and return water at a temperature of 70 oC
giving us a design temperature drop of 10oC.
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On Fig.18 we have marked the various pipework lengths and have labelled each run (A, B, C etc). We have also numbered each radiator and marked its emission in kilowatts in Table 8.
Flow Temp. 80°C Return Temp. 70°C Radiators Steel panel type,
Radiator
wall mounted. Commercially available outputs selected to
eet heat loss requirements.
m Outputs marked in kW. 1 = 1.6 kW 2 = 2.2 kW 3 = 2.8 kW 4 = 1.2 kW 5 = 1.8 kW 6 = 2.4 kW 7 = 2.0 kW
Boiler
8 = 2.6 kW
Fig. 18 Pipe Sizing Example Basic Information In order to use pipe sizing charts, we have to determine the mass flow rate of water which we need to supply to each radiator. Given radiator emission and temperature drop, the formula for this is given by Equation 11.
o
o
Where :
Mass flowrate = kg/s
Heat emission = kW
Temperature drop = C
S.H. of water = 4.19 kJ/kg C
Heat emissionMass flowrate = Equation 11
Temperature drop x S.H. of water
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For instance, for radiator 1, from Equation 11;
1.68Mass flowrate (kg/s) =
10 x 4.190.04 kg / s=
We can now populate Table 7, showing: a) Radiator emission b) Add a percentage for heat loss from the tail pipes serving each radiator c) The water requirement of each radiator
Note that item (b) needs to be carefully applied. This allowance will depend on the lengths of the tail pipes. It can normally be assumed that these pipes will be relatively short to each radiator, run indoors, and any heat loss from them will naturally be contained inside the building that is being heated. If this is so, an extra allowance of 5-10% of each radiator’s emission rate seems reasonable. Table 7 Converting Radiator Emission to the Water Requirement
Radiator number Radiator emission +5% heat losses from tail pipes
(kW)
Water mass flowrate
(kg/s) (kW)
1 1.6 1.68 0.040
2 2.2 2.31 0.055
3 2.8 2.94 0.070
4 1.2 1.26 0.030
5 1.8 1.89 0.045
6 2.4 2.52 0.060
7 2.0 2.10 0.050
8 2.6 2.73 0.065
We are now able to mark the mass flow rate figure against each radiator. Starting from the radiator furthest from the boiler, we can then progressively add up the loads to be carried by each section of pipe, right back to the boiler (Fig.18). For example; Pipe run K must carry sufficient water to serve:
Radiator No. 8 = 0.065 kg/s. Pipe run J, must carry sufficient to serve:
Radiators Nos. 7 and 8 = 0.065 + 0.050 = 0.115 kg/s . . . and so on.
Index circuit We are almost ready to size the pipework. But there is one more item we must determine before doing so. It is to decide which is the "Index Circuit". With a pumped system, this is the piping circuit which serves the heat emitter furthest from the boiler / pump (the Index Radiator).
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42
Looking at our system (Fig 18) and the lengths of pipework, the Index Radiator is No.8 and, thus, the Index Circuit is A, C, D, F, G, H, J, K. We will see later on that the resistance in the index circuit is used when the pump selection process is carried out. In simple terms, if we provide enough pump energy to deliver the right quantity of water to the furthest radiator, we will be able to satisfy the needs of all the other circuits and radiators. (Note: There can be exceptions to the "furthest radiator" rule. Where, for instance, a special piece of equipment is to be served which has an exceptionally high resistance to flow). Using the dimensions given on Fig.18, and the mass flow rates we have already established (Table 7) we are now able to construct a basic pipe sizing table for our example (Table 8). The index circuit and each sub-circuit are kept separate and the lengths of each pipe run are shown. These are the actual lengths and must include both the flow and return e.g. pipe run A, both flow and return is 2 x (2 + 5) = 14 m. It is also necessary, at this stage, to add a percentage to the actual length to cover pipe fittings and arrive at a "total equivalent length". In other words, we are converting the bends, tees etc into equivalent lengths (Ie) of straight pipe. Since the piping runs in the example are simple, we will allow a 10% margin. Table 8 Pipe Sizing Table – First Stage Section of pipe
Serving radiators
Water mass flowrate
(kg/s)
Actual length of flow & return
pipes (m)
+10% for fittings = le (m)
Index circuit
A 1,2,3,4,5,6,7,8 0.415 * 14 15.4
C 2,3,4,5,6,7,8 0.375 20 22.0
D 3,4,5,6,7,8 0.320 17 18.7
F 4,5,6,7,8 0.250 7 7.7
G 5,6,7,8 0.220 8 8.8
H 6,7,8 0.175 6 6.6
J 7,8 0.115 18 19.8
K 8 0.065 22 24.2
Sub circuits
B 1 0.040 16 17.6
E 3 0.070 10 11.0
I 6 0.060 14 15.4
* This is the flowrate the boiler and pump need to handle. Referring back to Table 5, under each pipe size a mass flowrate is listed together with a frictional resistance or pressure loss per unit length (Pa/m) on the left hand side. The only information we have in our example is the mass flow rate - we do not yet know the pressure loss or the pipe size.
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43
Putting this another way, we have one 'known' and two "unknowns'. So, in order to make use of Table 5, we must assume a value for one of the 'unknowns'. It is customary to assume a value for the pressure drop per unit length of between 200 to 450 Pa/m in order to complete our preliminary pipe sizing. Why not higher than 450 Pa/m? Experience has shown that if this upper limit is exceeded we may run into velocity noise problems and/or the pump necessary to provide the motive force will have a fairly high head with excessive power consumption. Why not lower than 200 Pa/m? This is a commercial issue. If a lower limit is adopted, pipe diameters one or two sizes larger may result with a consequent increase in capital cost etc. To generalise, the use of the 200 - 450 Pa/m pressure loss will ensure reasonable pipe sizes both from a noise aspect and from a commercial viewpoint. It will also ensure that the pump necessary for the job is well within the range available. We now know the mass flow rate for each pipe section and have assumed a pressure loss value. This enables us to use Table 5 to determine the actual size of pipe that we require. Initially, only the Index Circuit is considered. We know that pipe A must carry 0.415 kg/s. Using Table 5 and keeping the pressure loss between 200 and 450 Pa/m, we would select a 25mm diameter pipe. So Pipe Section A, carrying 0.415 kg/s will have a pressure loss, or frictional resistance of some 334 Pa/m. The Pipe Sizing Table can now be extended to include columns for pipe diameter and pressure loss / unit length. Multiplying the latter by the Total Equivalent Length will give us the Actual Resistance e.g. for Pipe Section A, 15.4 metres x 334 Pa/m = 5 144 Pa. Table 9 shows this extended detail for the Index Circuit.
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44
Table 9 Pipe Sizing Table – Second Stage Section of
pipe Index Circuit
Mass flowrate
(kg/s)
Actual length
(m)
Equivalent length
(m)
Pipe nb from Table 6
DN (mm)
Head loss from Table
6 (Pa/m)
Total head loss
(Pa)
A 0.415 14 15.4 25 334 5 144
C 0.375 20 22.0 25 275 6 050
D 0.32 17 18.7 25 204 3 815
F 0.25 7 7.7 20 406 3 126
G 0.22 8 8.8 20 320 2 816
H 0.175 6 6.6 20 207 1 366
J 0.115 * 18 19.8 20 95 1 881
K 0.065 * *
22 24.2 15 160 3 872
TOTAL 28 070
* In the case of pipe J, the mass flowrate of 0.115 kg/s is above 450 Pa/m for DN15 steel tube and below 200 Pa/m for DN20 steel tube. The better decision would normally be to select the larger pipe, and err on the side of caution * * In the case of pipe K, the mass flowrate of 0.065 kg/s is below 200 Pa/m for DN15 pipe, but this pipe is chosen as it is the smallest pipe in Table 5.
Sub-circuits It is now necessary to size the sub-circuit pipes B, E, and I. The procedure for this is slightly different to sizing the index circuit. Consider pipe section I, and referring to Table 8. Section I must carry 0.06 kg/s of water in order to serve radiator 6 with sufficient heat. The next step is to consider the pressure available to size Section I. The total actual resistance for the index circuit is 28 070 Pa. In getting to the point where sub-circuit I branches from the index circuit, we have used up the head losses of pipe sections A, C, D, F, G, H as shown in Table 10. Table 10 Pressure Loss Along Index Circuit A,C,D,F,G,H
Sections along the index circuit Pressure loss (Pa)
A 5 144
C 6 050
D 3 815
F 3 126
G 2 816
H 1 366
Total pressure loss 22 317
By subtracting this from the total index circuit, we arrive at the pressure head available to push the water through pipe section.
Hot Water Heating
Total index circuit pressure loss - index circuit A,C,D,F,G,H = 28 070 – 22 317 = 5 753 Pa We therefore have available 5 753 Pa pressure loss on which to size pipe section I, with its total equivalent length of 15.4 m and load of 0.06 kg/s. In order to use Table 5, this total pressure loss must be converted back to a pressure loss per unit length (Pa/m).
Thus, pressure loss, pipe section I = 5 753 Pa 15.4 m = 374 Pa/m
From Table 5, using 374 Pa/m as our basis for sizing, a DN15 pipe can be selected. This will be too large, and for a flowrate of 0.06 kg/s will produce a pressure drop of only 136 Pa/m. So, pipe section I, we will only use 136 Pa/m x 15.4 m = 2 095 Pa of the 5 753 Pa available to us. This leaves us a margin of 5 753 – 2 095 = 3 658 Pa ‘unused’. This will be dealt with more fully in a later section on ‘balancing’ but, in short, we have to consider absorbing this excess pressure loss by creating an artificial resistance via a Lockshield valve on the radiator. The same procedure is carried out for sub-circuits B and E, and Table 11 shows the completed results. Table 11 Pipe Sizing Example : Sub-Circuits Section of
pipe Index Circuit
Mass flowrate
Actual length
Equivalent length
Pipe nb from Table 6
Head loss from Table
6
Total head loss
(kg/s) (m) (m) DN (Pa) (Pa/m) (mm)
B 0.040 16 15 17.6 65 1 144
E 0.070 10 15 11.0 180 1 980
I 0.060 14 15 15.4 136 2 095
TOTAL 5 219
Fig.19 shows the pipe sizes we have selected.
45
Hot Water Heating
46
Preliminary Pipe Sizes A = 25mm
B = 15mm C = 25mm D = 25mm E = 15mm F = 25mm G = 20mm H = 20mm I = 15mm
J = 20mm
K = 15mm Boiler
Fig. 19 Pipe Sizing Example – Preliminary Pipe Sizes We would also carry out the sizing for the short pipes serving Radiators 2, 4, 5 and 7 in a similar manner to that used for sub-circuit sizing. The pipe sizing example has highlighted the fact that because we must use commercial pipe sizes, it is usually impossible to use up all the available pressure loss when sizing sub-circuits. An artificial frictional resistance must be created in each sub-circuit (some more than others) in order to absorb the available pressure loss and prevent "short-circuiting". This is the basis of the need for balancing which, as was stated before, will be dealt with later. The pipe sizing example may seem laborious but it does clearly show the areas where severe balancing will be necessary. It also enables the designer to try to use up all the available pressure loss when sizing sub-circuits although, in the example, commercial pipe sizes made this impossible. There are many methods for carrying out pipe sizing but the principles involved are the same. The pipe sizes we have determined would probably be adequate for a simple, small installation. However, on a large or complex system, what we have done so far would only be termed Preliminary Sizing as it would be necessary
Hot Water Heating
47
to carry out much more exact Final Sizing.
Having established preliminary pipe sizes, the object of final sizing is to eliminate the percentage allowances (at best an educated guess) which were made for heat emission from the piping and for the resistance of fittings.
Knowing an, albeit 'preliminary', pipe size, this can be used to determine the actual heat emission from each section of piping. Which leaves the allowance we made for pipe fittings. In Table 5, to the right of each mass flow rate column, there is a column marked ‘Ie’. This is the equivalent length of pipe in metres when K = 1.0. K is the velocity head or friction factor which varies with each type of fitting as can be seen from Table 6.
Thus the type and number of fittings in each section is noted and the velocity head factor K is multiplied by the Ie figure to give the additional length of pipe to be added to the actual length.
Let us imagine that we have 16 m of DN40 pipe carrying 1.053 kg/s of water. From Table 5, the pressure loss is 200 Pa/m and Ie = 1.8 m. The pipe is complete with two 90o bends and two 45o elbows.
Final pipe sizing K values of fittings DN40 90o bends K = 0.5 each (2 off) DN40 45o elbows K = 0.6 each (2 off) So the total K value for the fittings = (2 x 0.5) + (2 x 0.6) = 2.2 Le for the DN40 pipe = 1.8 m So, Equivalent length for the fittings = 2.2 x 1.8 m = 4 m Therefore, The whole equivalent length including that of the pipe itself = 16 m + 4 m = 20 m In our preliminary pipe sizing, we may have allowed only 10% for the fittings’ friction loss, giving us a total equivalent length of 16 m + 10% = 17.2 m. The K values for fittings allows us to calculate more accurately the piping resistance, and some revisions to the preliminary pipe sizing results may be necessary. Proportioning mains losses One last important part of final pipe sizing which should be mentioned is ‘proportioning’. To explain this, let us take the situation shown in Fig. 20 as a very simple example.
Hot Water Heating
4000W Flow Rate = 0.096 kg/s
3600W Flow Rate =0.086 kg/s
3000W Flow Rate = 0.07 kg/s
A B C
1 2Flow 80oC
Return 70oC Mains loss = 410W = 0.01 kg/s
Mains loss = 380W = 0.009 kg/s
Mains loss = 240W = 0.006 kg/s
Fig. 20 Proportioning Mains Losses Consider radiator C Flow rate needed for mains loss in section 2 = 0.006 kg/s Flow rate needed for radiator emission = 0.070 kg/sTotal required flowrate (for section 2) = 0.076 kg/s Consider radiator B Flow rate needed for mains loss in section 1 = 0.009 kg/s Flow rate needed for radiator emission = 0.086 kg/sTotal required flowrate (for section 1) = 0.095 kg/s But the 0.009 kg/s is the flowrate necessary to meet the mains loss in pipe section 1. To counteract this loss, radiators B and C must be supplied with this extra load in proportion to their own flowrates needed for their design emissions. Therefore,
0.009 kg/s x 0.076 kg/sRadiator C needs to take up an extra = 0.004 kg/s
(0.086 kg/s + 0.070 kg/s)
0.009 kg/s x 0.086 kg/sRadiator B needs to take up an extra = 0.005 kg/s
(0.086 kg/s + 0.070 kg/s)
After accounting for these proportional ‘extras’ needed for mains losses, the adjusted total flowrates become:
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Hot Water Heating
Pipe section 2 (rad C + mains loss) = 0.076 kg/s
0.004 kg/s
0.080 kg/s
Pipe section 1 (rad B + mains loss) = 0.086 kg/s
0.005 kg/s
0.091 kg/s
+
+
If we did not supply this "extra" water to take into account heat lost from the mains, then radiators B and C would be supplied with water below the design flow and return temperatures. It is obviously more accurate than our original preliminary allowance of adding a percentage to each radiator output to cover the heat emission from the mains, since we have now not only calculated the actual heat loss but have "shared" this among the heat emitters in proportion to their requirements. To explain this another way, it is necessary for the most distant radiators to be maintained at the same mean temperature as those nearest to the boiler. But the heat lost from the flow main will have cooled the water entering those radiators at a distance, more than those nearer the heat source. Similarly, water leaving these radiators will be cooled still more in the return main back to the boiler. The only way in which the mean temperature can be kept uniform throughout, is for the most distant radiators to pass more water than the nearer ones. Or, in other words, that they should be selected based on a smaller temperature drop. Proportioning the mains loss achieves this objective without calculation of actual temperature drops. It can be a tedious process on large jobs and is quite often conveniently overlooked. This can result in heat starvation since insufficient heat may be carried by the pipes to meet required duties. Several alternative methods of carrying out proportioning are available both graphically and arithmetically. The truly optimistic footnote to the whole subject of pipe sizing is that much of it is now carried out via computer programs.
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Hot Water Heating
SELECTING THE BOILER AND PUMP Having carried out the heat loss calculations, selected and sized the heat emitters and the piping system to connect them, it is now necessary to select the boiler and pump to heat and circulate the water around the system.
Boiler heat capacity In order to select the correct boiler, we need to establish the required boiler heat capacity. This is done by using Equation 12.
o
o
.
p
.
p
Boiler capacity (kW) = m c T Equation 12
Where:
Δ
Boiler capacity in kW
m = Mass flowrate (kg/s)
c Specific heat of water (kJ/kg C)
T = Temperature difference between boiler flow and return ( C)
=
Δ
p
Boiler capacity (kW) = m c T Equation 12
m = 0.415 kg/s
c 4.19 kJ/kg C
T = 10 C
Boiler capacity = 0.415 x 4.19 x 10
= 17.4 kW
Δ
=
Δ
Returning to our pipe sizing example (Example 3), the mass flowrate to and from the boiler would be 0.415 kg/s (from Table 8), and the temperature difference between the boiler flow and return water upon which the system is designed is 10oC (80°C flow and 70°C return). From Equation 12:
o
o
.p
.
Where:
It would be normal practice to add a safety capacity margin to allow for perhaps a domestic hot water requirement via, perhaps, an indirect heating cylinder, or to cater for the inevitable brief periods in winter when the outside temperature falls below the considered design temperature. Also, it is unlikely that a boiler of exactly the right size would be commercially available; we would be forced to accept a boiler from a manufacturer’s range. From the above, we may select a boiler with, perhaps, a 24 kW capacity rating. Large volume buildings with high ceilings with intermittent heating needs (such as a church, perhaps heated only once a week) would be treated quite differently, and is outside the scope of this document.
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Hot Water Heating
Varying the flow or return temperature Our whole system in Example 3 has been based on design parameters of 80°C flow and 70°C return water temperatures; that is, on a 10°C temperature drop across the radiator. This basic fact was used to establish the heating areas of the radiators to give the required heat outputs, and to size the distribution pipework with acceptable water velocities and friction losses. A definite relationship exists between the radiator flow & return temperature drop, the radiator heating area, and the water flowrate through the radiator. Changing the design temperature drop will always alter the water flowrate, but not necessarily the radiator size. Consider the following two examples, both for a radiator heat loss of 4 kW.
Example 4 Increasing the temperature drop, but reducing the mean temperature difference between the radiator and room. What would be the situation if the flow temperature were the same, but the temperature drop across the radiator were increased from 10°C to 20°C, that is, with an 80°C flow and 60°C return temperature? Let us assume that the design room temperature remains at 20°C in both examples. In the initial case (80°C flow/70°C return; mean=75°C), the mean temperature difference between the radiator and the room is 75°C - 20°C = 55°C. In the second case (80°C flow/60°C return; mean=70°C), mean temperature difference between the radiator and the room is 70°C - 20°C = 50°C. Because of the lowering of the differential temperature between the radiator and the room from 55°C to 50°C, the radiator heating area would need to be greater to compensate in the second case. However, there is some benefit to be gained from the second case. As the temperature drop increases from 10°C to 20°C, the water flowrate will fall in inverse proportion, as depicted by Equation 11. For the first case, with a temperature drop of 10°C;
o o
Heat emissionMass flowrate = Equation 11
Temperature drop x S.H. of water
4 kWMass flowrate =
10 C x 4.19kJ/kg C
Mass flowrate 0.096 kg / s=
For the second case, with a temperature drop of 20°C;
o o
4 kWMass flowrate =
20 C x 4.19kJ/kg C
Mass flowrate 0.048 kg / s=
As the second case temperature drop is twice the first, the flowrate is halved. Fig. 21 sums up the effect.
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Hot Water Heating
Space Temperature 20oC
Mean Water Temperature
75oC
a) Temperature Difference (75 – 20) = 55OC
80oC 70oC
Space Temperature 20oC
80oC 60oC
Mean Water Temperature
70oC
b) Temperature Difference (70 – 20) = 50OC
i.e. heating surface in (b) must increase by 10% to give the same output as (a)
Fig. 21 Effect on radiator heating area by reducing the radiator/room mean temperature difference So, a higher temperature drop between the flow and return temperatures will always result in lower water flowrates, which can mean smaller pipes and lower capital cost. A smaller pump duty will be needed, resulting in lower running costs. However, a higher flow and return temperature drop does not always mean a larger radiator. If the difference between the mean radiator temperature and the room temperature remains the same, the radiator remains the same size. This can be achieved by using a higher flow temperature but a lower return temperature. Compare Example 5 to Example 4.
Example 5 Increasing the temperature drop by increasing the flow temperature, but retaining the mean temperature difference between the radiator and room. Let us assume that we design the system on a higher flow temperature of 90°C and a return temperature of 60°C. In the initial case (80°C flow/70°C return; mean=75°C), the mean temperature difference between the radiator and the room is 75°C - 20°C = 55°C. In the second case (90°C flow/60°C return; mean=75°C), mean temperature difference between the radiator and the room is also 75°C - 20°C = 55°C. We can see that the mean temperature difference is the same for both situations, therefore no increase in heating area is necessary to give the same required heat output.
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Hot Water Heating
53
Flow and return design temperatures tend to change depending on national preference. For instance, in the UK, LTHW radiator temperature drops have traditionally been 10°C and, latterly, 20°C; whereas in mainland Europe, temperature drops of 40°C are often preferred by designers.
Hot Water Heating
54
Pumps and pump sizing Another major item that needs to be selected for our simple system to work is the circulation pump, correctly sized with the right duty. From Table 8, we know that the total mass flowrate of water that we must move around the system is 0.415 kg/s. We also know from Table 9, that the total pressure loss (frictional resistance) of the index circuit is 28 070 Pa. In essence, this is the duty that the pump must provide. However, pumps are normally rated on a volumetric basis. Converting from mass flowrate to volumetric flowrate is a simple matter, as the density of water only varies by about 3.5% between 10°C and 90°C, and it is usual to consider that a litre of water weighs 1 kg for LTHW systems. It is also usual to express the pump head in terms of kilopascals (kPa), due to the large magnitude of numbers involved. For Example 3, we would start by selecting a pump to handle 0.45 kg/s against a frictional resistance of 28.1 kPa from a commercially available range. It is important that a pump is selected with a duty as near as possible to our requirement; as we shall see shortly. In other words, the addition of large ‘safety’ margins should be avoided. Important though all the components of a heating system are, the pump could be considered as being the ‘heart’ of the system. Incorrect selection will affect the successful operation of the whole system. The way that the system is controlled and operated will affect the hydraulic properties of the system and the effect of the pump. It seems appropriate then, to spend some time covering further aspects of pumps and pumping in relation to our hot water heating system. A pump is a machine that adds mechanical energy to a fluid for the purpose of moving the fluid from one point to another. This ‘addition of energy’ is usually achieved by one of two methods; by using a reciprocating pump, that is, a pump having a piston and cylinder; and a rotary pump, such as a centrifugal pump, which has a rotating mechanism to move the fluid. Since a centrifugal pump is usually the type used in LTHW heating systems, we will concentrate on them.
Centrifugal pumps A centrifugal pump comprises an impeller (a rotating disc equipped with suitably shaped vanes) rotating at high speed within a stationary casing. The impeller is driven by an electric motor. The incoming liquid is directed into the centre or ‘eye’ of the impeller, from where, due to its fast rotation, the liquid is thrown out, via the vanes, to the impeller edge at high velocity. This causes the kinetic energy of the liquid to have increased substantially at the point where it leaves the impeller. The liquid enters the pump casing which is duly shaped to offer an increasing volume to the liquid flow as it passes through it. The increasing volume causes the liquid velocity to drop and, in accordance with Bernoulli’s principle, the falling kinetic energy changes into pressure energy, the highest pressure being at the pump outlet. Also, on the inlet side of the pump, as the liquid leaves the eye of the impeller, it creates a pressure drop (or suction), which sucks in liquid to replace that which has passed through the pump.
Hot Water Heating
The pump therefore creates a situation where its inlet pressure is reduced and its outlet pressure is increased. The nett effect is that a pressure drop is created over the pump, which causes a circulation of liquid through it. The advantages of a centrifugal pump for use on heating systems is that they operate at fairly high speed. This results in acceptable size and cost; a continuous delivery free from pressure fluctuations; and can be conveniently and simply driven by an electric motor and operate with predictable characteristics. Centrifugal pumps can have one impeller and are called single-stage pumps; can be multi-stage with several impellers; direct coupled to an electric motor via a flexible mechanical coupling; belt driven or close-coupled via a common shaft for both motor and impeller. Most pumps used in contemporary LTHW systems are of the canned rotor (or wet rotor) type, where the rotor is designed to allow its rotor to run in water. The stator is protected from water by encapsulating the rotor in a ‘can’ enclosure (Fig 22), or by other means. These pumps can be available with either fixed or variable pressure output.
55
Fig. 22 Canned Rotor Pump Varying the amount of water allowed to spill around the edge of the impeller will vary the pressure or ‘head’ developed by the pump. This can be achieved either by an internal valve mechanism controlling a path between the inlet and outlet sides of the impeller, or by means of varying the clearance between the impeller and the body or casing of the pump.
Sta
tor
Sta
tor
Rot
or
Rot
or
Pump characteristics Pump performance curves are produced by pump manufacturers to allow designers to choose the correct pump for their system. Such curves usually show the relationship between the pump outlet pressure, the flow capacity, and the power consumed and efficiency. Typical curves are depicted in Fig. 23.
Hot Water Heating
Fig. 23 Typical pump curves for a Constant Speed Centrifugal Pump
Pressure/Capacity
Efficiency
Power (BHP)
Pre
ssur
e H
Hor
sepo
wer
Effic
ienc
y
Capacity Q
As can be seen from Fig.23, performance falls from a maximum pressure with little or no flow, to the reduced pressure which would be produced under maximum flow conditions. The pressure/capacity curve may be either steep or flat as shown in Fig.24 depending on the particular pump chosen.
Pre
ssur
e H
Steep
Flat
Capacity Q
Fig. 24 Comparing steep and flat performance curves Pumps with steep curves are frequently selected for ‘open’ systems, such as cooling systems with cooling towers. They have the advantage, in this case, that a change in piping frictional resistance, resulting from corrosion or scale accumulation over the lifetime of the installation, has a minimum effect on flow. However, for ‘closed’ systems, flat curve pumps are more often used because large changes in pump capacity can be achieved with a small change in
56
Hot Water Heating
pressure – this can be very useful when faced with having to balance a number of systems. The particular type of pump will, therefore, depend on the particular requirements of the circuit.
Pump laws Certain physical laws govern the relative performance of centrifugal pumps, which can be stated briefly as follows:
1. Pump flowrate varies directly as the speed of rotation. .
2 2.
11
m NNm
=
2. Pressure developed varies as the square of the speed of rotation. 2
2 2
1 1
H NH N
⎡ ⎤= ⎢ ⎥⎣ ⎦
3. Power absorbed varies as the cube of the speed of rotation. 3
2 2
1 1
P NP N
⎡ ⎤= ⎢ ⎥⎣ ⎦
Where: .
m = Pumping rate N = Speed of rotation of impeller H = Pressure P = Power absorbed by the pump
Pumps fitted in series As can be seen from Fig. 25, with two identical pumps connected in series, the pressure developed is twice that developed by a single pump but the capacity remains the same. The power absorbed will also be twice that of a single pump.
Pre
ssur
e H
Pumps in Series
Single Pump
Capacity Q Fig. 25 Pumps in Series
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Hot Water Heating
Pumps fitted in parallel With two identical pumps operating in parallel, as shown in Fig. 26, the capacity delivered by the two pumps is doubled that of a single pump, however, the pressure developed remains the same. The overall power absorbed by the pumping station will be twice that of a single pump.
Pre
ssur
e H
Pumps in Parallel
Single Pump
Capacity Q Fig. 26 Pumps in Parallel
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Hot Water Heating
SYSTEM CHARACTERISTICS Before it is finally possible to select a pump, it is necessary to consider the circulation system into which the pump will be fitted. Each system will have its own hydraulic characteristics concerning how much pressure is needed to achieve the required capacity. The relationship between pump pressure and capacity can be plotted on a graph as the system performance curve. A centrifugal pump responds to the hydraulic characteristics of the system to which it is applied and produces water flow and head (pressure) related to the pressure conditions of the system for a specific flow. This makes it necessary to carefully predict the system characteristics for proper pump selection. In the piping system, there is resistance to water flow because of pipe friction as seen in the previous sizing examples. This frictional resistance will vary with the amount of flow. If we push more water through a given size of pipe, the velocity will increase and, as a result, so will the frictional resistance. The friction increases with the amount of flow in almost exact proportion to the square of the velocity of the water and, for the same sized pipe, the velocity is proportional to the flowrate. This means that, if the water flow is doubled, the water velocity is doubled, but the frictional resistance increases fourfold.
The system curve This physical relationship between pressure and water flow can be expressed in simple arithmetic by Equations 13 and 14.
2.
.
P m Equation 13
Where :
P Pressure loss due to friction
m Flowrate around the circuit
α
=
=
Equation 14 is derived from Equation 13 in as much that the pressure loss at any two flowrates is proportional to the ratio of the flowrates.
2.11
.2
2
.11
.22
P m = Equation 14
P m
Where :
P Pressure loss due to friction at flow m
P Pressure loss due to friction at flow m
⎡ ⎤⎢ ⎥⎢ ⎥⎣ ⎦
=
=
An example should clarify any doubts for which we will use the figures obtained in Example 4, the pipe sizing exercise. The rate at which we want to circulate the water is 0.415 kg/s.
.1So, m = 0.415 kg/s
From Table 9, the index circuit pressure loss is 28.1 kPa. 1So, P = 28.1 kPa
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Hot Water Heating
In order to develop a system curve, we must now assume various values for and, from Equation 14, we can calculate and, in Table 12, record the
corresponding values for P
.2m
2.
Table 12 - The Values of for Different Values of P.
2m 2
1.6 5.8 13.4 18.6 28.1 35.9 P2 (kPa)
.2m (l/s) 0.1 0.2 0.3 0.35 0.415 0.45
A graph can now be constructed from these data, as shown in Fig. 27.
Pre
ssur
e P
System Curve
Capacity Q
Fig. 27 Typical System Performance Curve
The pump curve On the graph in Fig. 27, we can superimpose the pump performance curve (which would be supplied by the pump manufacturer) for the pump which has been selected for our pipe sizing example (Example 3), as shown in Fig. 28.
Pre
ssur
e P
System Curve
Pump Curve
32.5
Point of Pump Operation
0.476
Capacity Q Fig. 28 System Curve for Pipe Sizing Example
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Hot Water Heating
The point of intersection of the system curve and pump curve will be the point of pump operation; the flowrate and pump output pressure when coupled to the system. Because we have sized the pipe accurately and selected the pump carefully, the pump performance matches the system parameters closely, as can be seen from Fig. 28. But this is not always the case! A number of factors can cause pump characteristics and system performance curves to mismatch, including:
1. Errors in pipe sizing 2. Incorrect pump selection 3. Arbitrary adjustment of the frictional resistance figures
Assume that point 3 is causing a mismatch, and that a margin has been added to the calculated frictional resistance figures. The actual pressure losses in the system will be lower than the adjusted figure. Also, the pump has been selected against the selected figure. The actual system curve will be lower than predicted and, consequently, the actual pump operating point will shift, as shown in Fig. 29. This can cause problems in balancing the system and unnecessary waste in increased power used by the pump. Of course, the reverse of the above is that the frictional resistance of the system is greater than that calculated. In this case, the result will be that a lower flowrate of water will be circulated than that needed, causing ‘heat starvation’ to some of the radiators. The effects of temperature controls on both system curves and pump performance must not be overlooked. These have a marked influence on the hydraulic dynamic of the system, and this will be dealt with later.
Pump Curve
Pre
ssur
e P
Predicted System Curve
Predicted Operating Point
Actual System Curve
Actual Operating Point
Actual pressure loss in system at predicted flow
Capacity Q Fig. 29 Effect on Pump Performance of Difference between Predicted and Actual System Curve Table 13 is a simplified summary showing the various relative effects on centrifugal pump performance if the hydraulic conditions are altered.
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Hot Water Heating
62
Table 13 Effects of Various Condition Alterations to Pump Characteristics
Alteration Condition Result
Pump response to throttling discharge
Friction losses Power
Increase Decrease
Result of increasing the friction losses
Flowrate Power
Decrease Decrease
Result of decreasing the friction losses
Flowrate Power
Increase Increase
Hot Water Heating
POSITION OF PUMP IN SYSTEM
Introduction The object of a circulating pump is to impart mechanical energy to the water to cause it to circulate around the system via the piping network. The pressure energy created by the pump is spent in overcoming the frictional resistance which exists between the flowing water and the internal pipe walls. As far as supplying this energy is concerned, it makes no difference where the pump is fitted within the system. If the pipes and pump have been sized and selected correctly, and the system correctly balanced, the energy imparted by the pump will have all been dissipated by the time the water returns to the pump to be re-energised. Fig. 30 shows some alternative positions where the pump might be installed.
Alternative Pump Positions A, B or C
Feed & Expansion Tank or
Pressurisation Unit
Flow
A
Heating System
ReturnB C
Fig. 30 Alternative Pump Positions However, it is necessary to consider the alternative pump positions, and their effects on the distribution of pump pressure. Let us begin with a simple U-tube, as depicted in Fig. 31, into which we pour some water. The water will find its own level and be the same in columns A and B since atmospheric pressure is acting equally on each water surface in each column. As there is no movement in the water, the atmospheric pressure can be considered as imparting static pressure or static head to both water surfaces.
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Hot Water Heating
Static Pressure
Static Water LevelA B
Fig. 31 Simple U-Tube In Fig. 32, a pump is installed at the bottom of the connection between columns A and B.
X A B
Static Water Level
X
Fig. 32 U-Tube with Pump When energised, the pump will pull down the water in column B and push it up in column A. As long as the pump is installed in the middle of the two columns with exactly the same static head on either side, and the pipes are the same size, the fall in water level in B will equal the rise in water level in A. If we consider that the water level falls a distance ‘X’ in column B, then the rise in water level in the column A will also be ‘X’. This means that the pump is generating a pressure equal to the addition of the two distances (or heads), so the total pump head is 2X. For simplicity Fig. 33 shows that we now have fitted a header tank to the top of column B, though a pressurisation unit would of course have the same effect.
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Hot Water Heating
A
2X
Static Water Level
B
Fig. 33 U-Tube with Pump and Header Tank When the pump is energised, the volume of water held above the static water line in column A will be exactly the same as that held in the fall in the tank water level. If we consider that the volume of water held in the tank is very much more than the volume held in column A, then we can assume that, when the pump is energised, the fall in water level in the tank will be negligible, perhaps very close to zero. We could therefore say that, if the fall in tank water level is zero, the rise in column A is 2X. The next step is to build the header tank principle into a typical heating system, perhaps similar to that shown in Fig. 34. When the pump is energised, a small amount of the pressure generated by the pump moves the water up column A slightly, and the remainder of the pump energy is used to overcome the frictional resistance throughout the circulation system. Because of the extra volume of water held in the header tank, the water level in Column B does not fall.
Column A
ColumnB
O
Fig. 34 U-Tube System
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Hot Water Heating
As there is no flow through column B, the point ‘O’ shown in Fig.34 is subjected to static pressure only; any difference in pressure at point ‘O’ above or below static pressure being released by a re-balancing of the U-tube. Point ‘O’ is known as the ‘neutral point’ of the circulation system. The orthodox LTHW system can be developed from this principle, and a typical system is depicted in Fig.35. The column A is, in practice, directed over and into the header tank (but not underneath the water surface). This has two purposes; the first being to allow the water level to rise and fall in the pipe, depending on whether the pump is energised or not; the second, being to act as an expansion relief mechanism, to allow hot water to expand out of the system, so that the pressure cannot build up to dangerous levels. The header tank will also have an overflow, to allow any excess water to drain to a safe point. Because the header tank has this dual function, it is often referred to as a ‘feed and expansion tank’.
Feed & Expansion Tank
Heating SystemCold
Feed
Pump
Boiler OPoint of Neutral Pressure
Fig. 35 Point of Neutral Pressure in LTHW Systems Developing on from this, we need to consider the relative merits of alternative pump positions. Let us first consider a very simple circulation system consisting of one circuit only, and the piping being all at the same level around the loop, from and to the boiler.
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Hot Water Heating
A
B
C
Fig. 36 Simple LTHW Circuit as Pump Position example The static head (the pressure due to the height of water in the feed and expansion tank above the level pipework) will be constant in the sections A-B, B-C, and C-A. Wherever the pump is fitted in these sections, gauges fitted at the inlet and outlet sides of the pump will show the same ‘change in pressure’ while the pump is running, and the same static pressure when the pump is not running. The actual pressure recorded on the gauges while the pump is running will vary according to the position in the circuit where the pump is fitted, but the differential between the actual pressures recorded on the two gauges will be constant. The change in the water pressure, due to its passage through the pump, can appear as an increase in pressure at the pump outlet above the neutral point static head, a reduction in pressure at the pump inlet below the neutral point static head, or a combination of both. There is one point in the circulation system where the pressure cannot change, this being the point at where the cold feed is connected, that is, the neutral point. The pressure in the system at this point is supporting the column of water in the feedtank which is, in turn, providing the static head to the circulation system. (This can be somewhat likened to the neutral connection in a single phase electrical circuit, which is formed from the common ‘star point’ connection on the low voltage secondary side of a three phase delta-star electrical transformer). Pump fitted at point A Consider the pump fitted at point A in Fig. 36. The flow through the circuit is anti-clockwise, that is, from A to B to C to A. The piping from the pump outlet to point C is subjected to some increase in pressure above the neutral point, whilst the section between point C and point A will experience a pressure lower than the neutral point. The lowest pressure being recorded at the pump inlet. Fig. 37 shows how the greatest pressure is formed at the pump outlet at A, and how the pressure drops as it passes through the circulation system, due to its losing energy by friction from the pipe wall. The pressure equals the static head at the neutral point (just after point C), and continues to fall a little
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Hot Water Heating
more in passing through the short run of pipe to, through, and from the boiler, and on to the pump inlet, where it is at its lowest.
O = Static Pressure
OA B C
Fig. 37 Pump at Point A Because, by far, the greatest part of the circulation system is under positive pressure, we can say the system is a ‘positive pressure system’. Pump fitted at point C If the pump were fitted at point C, the neutral point would be close to the pump outlet, and as the neutral point is exclusively tied to the static head pressure, the pump cannot increase the pressure at its outlet above the neutral point. As far as the pump is concerned, it does not care or know if its outlet pressure is higher or equal to the neutral point static head; it only exists to provide a relative positive pressure difference between its inlet and outlet ports. As the pump outlet pressure is tied to the static head pressure at the neutral point, it produces its pressure differential by having a low pressure at its inlet port. As in the previous situation, the lowest pressure in the circulation system is recorded at the pump inlet. A B C
O
O = Static Pressure
Fig. 38 Pump at Point C Fig. 38 depicts the pressure drop around the system. We can see that, in this instance, because the pump outlet pressure is the same as the static head, due to its proximity to the neutral point, all other pressures around the circuit fall in the direction of flow; the lowest being at the pump inlet. As all pressures in the circulation system are relatively lower than the static head, we can think of this system as operating with a negative pressure, and we can term this system a ‘negative pressure system’. Pump fitted at point B
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Hot Water Heating
If the pump were fitted at point B, some of the circuit would be under positive pressure and some under negative pressure, relative to the neutral point. This situation is depicted in Fig. 39.
A O
B C
O = Static Pressure
Fig. 39 Pump at Point B
Summary of pump positions So the pressure at the neutral point is always the static pressure irrespective of the pump position, so this can be used as a datum pressure to which all other circulation pressures relate. All that change are the relative pressures around the system, depending on where the pump is placed in the circuit. The distribution of the pump pressure around the circuit can be related to the relative position of the pump and the cold feed connection (the neutral point). This relationship is depicted in Figures 40, 41, and 42, indicating three alternative but commonly used situations. Positive signs are used to indicate that the circulation pressure is higher than the datum of the neutral point, whilst negative signs indicate a lower pressure.
Pump in return line, cold feed on SUCTION side of pump In Fig. 40, the cold feed supply is situated on the suction side of the pump, and both are fitted in the return line to the boiler. All pressures after the pump outlet are higher than the neutral point datum. The only pressures lower than the neutral point are those in the pipe between the neutral point and the pump inlet. As the distance between the neutral point and the pump inlet is relatively small, the majority of the system is subjected to positive pressure when the pump is running. Air removal (via cocks and air eliminators) from the piping system and radiators is easily facilitated.
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Vent
Rise in Water Level
70
Fig. 40 Pump in Return, Cold Feed on Suction, Vent on Flow Line In the system depicted in Fig. 40, the main disadvantage is that the open vent pipe must be carried up to a height in excess of the residual pump head above the level of the tank, to prevent water from continuously discharging from it. It is not always possible or practical to run open vents to the necessary height in cramped roof spaces.
Pump in return line, cold feed on DISCHARGE side of pump In Fig. 41, the pump outlet is situated close to the cold feed datum point and, as such, the only positive pressure in the system can be between these two points.
Boiler Pump
Flow
Return
Datum Line
PumpHead
Feed & Expansion Tank
Pump Pressure Distribution
Vent
Datum Line
Fall in Water Level
PumpHead
Feed & Expansion Tank
Pump Pressure Distribution
Flow
Return
Boiler Pump Fig. 41 Pump in Return, Cold Feed on Discharge, Vent on Flow Line Hence, most of the system operates below the static pressure when the pump is running, that is, at a negative pressure relative to the neutral point. Care must be taken to ensure that high level apparatus (including air eliminators) do not operate at sub-atmospheric pressure enabling air to be drawn into the
Hot Water Heating
system via valve glands etc. The only advantage of this system is that the vent pipe does not need to be extended very much above the level of the header tank, as the pressure in the vent pipe will always be lower than the static pressure when the pump is running, under normal operating conditions.
Pump in flow line, cold feed on SUCTION side of pump The system shown in Fig. 42 has all the advantages of the system in Fig. 40 but, additionally, it does not require the open vent to be taken above the level of the tank. This, again, is due to the fact that the water level in the vent pipe when the pump is running under normal conditions will be lower than the water level in the tank.
Vent Pump Pressure Distribution
Datum Line Feed & Expansion Tank
Fall in Water Level
Flow
Boiler
Pump
Return
Fig. 42 Pump in Flow, Cold Feed on Suction, Vent on Flow Line It is, however, possible for air to be drawn in through the open vent if the height of the vent pipe above the header tank is less than the circulation pressure at the vent pipe connection point. To generalise, the arrangement shown in Fig.42, with the pump in the flow, and the cold feed (and open vent) on the suction side, tends to be the one most widely favoured by designers of LTHW systems. It must just be mentioned that modern systems are designed without feed and expansion tanks. The feed supply is fed directly from the mains (via a Pressurisation Unit, as shown earlier) at a regulated pressure instead of relying on a tank providing the static head. The neutral point is still relevant and is still situated at the cold feed connection point.
Inverter Driven Pumps With ever increasing pressure on the reduction of energy, many pumps nowadays are fitted with inverter drives. The objective of this is for the pump output to be controlled to meet the required load of the system, rather than it running at constant speed. This is achieved by use of an inverter drive with associated sensor applicable to the application. Lifecycle cost analysis shows that only 5% of the lifetime cost of a fixed speed pump is initial capital and maintenance cost. The other 95% being consumed energy. Using variable speed drives to control the speed of a pump to match
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demand allows significant energy savings – a 20% reduction in speed equates to a 50% reduction in power absorbed.
Balancing the system During the pipe sizing example (Example 4), the need for proper balancing has been mentioned several times. The need is to absorb all the available pump pressure for each sub-circuit. Since commercial pipe sizes are unlikely to allow us to achieve this by accurate pipe size selection, extra resistance to flow is achieved by fitting regulating valves in strategic places which can be adjusted to ‘fine tune’ the system frictional losses to the pump pressure. Setting can be achieved by two methods.
Balancing by using temperature In a simple system which may, for example, have been designed for an 80oC flow and 70oC return, we may use temperature as the basis for setting the regulator valves. The quantity of water we need to circulate to meet the heat requirements is based on the 10oC temperature difference between the flow and return lines. So, if we adjust the temperatures on the return lines from each sub-circuit (or radiator) so that the return temperature is 70oC, the correct quantity of water is being circulated and the system is balanced.
Balancing by using pressure In a large circulation system, it is usually more suitable to fit a special balancing valve to each sub-circuit. This type of valve will have two pressure tappings fixed to its body, one upstream of the valve orifice and one downstream. From flow charts, the valve can be adjusted to give the required extra resistance to flow. We will have determined the required resistance for each sub-circuit during the pipe sizing calculations.
Fig. 43 Typical Balancing Valve and Differential Pressure Gauge
72
If the excess pump pressure available at each sub-circuit is not used up artificially (because the pressure loss is too low), the flowrate of circulated water will increase. In extreme cases, the ultimate point may be reached where the return temperature is almost the same as the flow temperature. This means that the water is circulating too fast to give up its heat to the
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73
radiators, and it has to be slowed down to desired levels by restricting the flow by a balancing valve. Conversely, a sub-circuit at the far end of the system may not have any circulation, since all the energy is being used up in passing the water through those circuits nearer the pump. To illustrate this further, Fig.44 shows an example of a larger system.
Hot Water Heating
2 kg/s 60kPa 3 kg/s 40kPa 5 kg/s 70kPa 1 2 3
Pump
30 kPa
50 kPa Boiler
Fig. 44 Balancing a Larger System Fig. 44 shows a large district heating system serving three separate buildings via the same distribution main. It can be seen from Fig. 44 that the total distribution friction loss is 80 kPa, and the largest sub-circuit friction loss is from the Building 3 sub-circuit, with 70 kPa. Therefore: Total distribution friction loss = 80 kPa Largest sub-circuit friction loss = 70 kPaPump pressure required = 150 kPa It can also be seen from Fig. 44 that the pump has to provide a flowrate of 10 kg/s. Building 1, which incurs a frictional pressure drop of 60 kPa, will be subjected to the full pump pressure of 150 kPa, as it offers the first sub-circuit resistance from the pump. The amount of water pumped into Building 1 sub-circuit can be calculated from Equation 14.
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Hot Water Heating
2.11
.2
2
1
2
.1
.2
P m = Equation 14
P m
Where :
P Required pump pressure (kPa)
P Friction pressure loss in circuit (kPa)
m Theoretical pump flowrate (kg/s)
m Design pump flowrate (kg/s)
Therefore;
150 kPa Y60 kPa
⎡ ⎤⎢ ⎥⎢ ⎥⎣ ⎦
=
=
=
=
=
2 kg/s2 kg/s
150Y kg/s = 2 kg/s x
60Y kg/s = 2 kg/s x 1.58
Y kg/s = 3.16 kg/s
⎡ ⎤⎢ ⎥⎣ ⎦
So, instead of receiving a design flowrate of 2 kg/s, the Building 1 sub-circuit will receive 3.16 kg/s. The same thing will occur with Building 2, and the result is that Building 3 will not receive enough flowrate, due to, what is essentially, a short-circuit on the first part of the system. To overcome this, balancing valves can be installed in the Buildings 1 and 2 sub-circuits, to create an artificial resistance in these circuits. This will reduce the flowrates to Buildings 1 and 2, to allow the required amount to circulate around Building 3 sub-circuit.
Summary of balancing This is a somewhat simplified view of balancing, but all too often it is the only attention paid to the subject – even if balancing is considered at all! Returning briefly to the subject of balancing on the return temperature, this will only be successful when the outside temperature is at design conditions. If temperature controls are installed to vary the water temperatures relative to the outside temperature, this method becomes unusable. Even the ‘pressure loss’ balancing method has its shortcomings. It assumes that a constant volume of water is always circulating around the system. As we shall later see, variable volume systems are not uncommon in commercial premises, and the effect of varying the flowrate will be to vary the hydraulic dynamic of the whole system. Automatic balancing techniques are best used to deal with these problems – but more on this subject when variable volume systems are discussed.
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76
AUTOMATIC CONTROL OF A HEATING SYSTEM This document is not intended to cover the wide-ranging subject of automatic control. But it is important that we deal with the basic methods of controlling the temperature of the building together with the effects of the controls on the building itself.
Let us return to the beginning. Our heating system has been designed to provide a ‘comfortable’ internal temperature (perhaps 20oC) against a worst average external temperature (perhaps -5oC). The system and heat emitters have been designed and selected to do this when supplied with hot water at specific flow and return temperatures (perhaps 80oC flow and 70oC return).
For most of the heating season, outside temperatures will be above -5oC, so if we are to prevent overheating with consequent uncomfortably high internal temperatures and energy wastage, we must find some way to control the amount of heat the system provides according to the actual (and varying) conditions.
What we are effectively saying is that, when the external temperature is above -5oC (in our example), the heating system is oversized, and will give out more heat than we need. We require some from of control to ‘derate’ the system’s heat output.
Only when the external temperature is at -5oC will the system be correctly sized to meet the heat loss from the building without any reduction in heat output from the heat emitters (assuming the calculations have been carried out correctly, of course).
If the outside temperature falls below -5oC, it will be difficult for the system to meet the desired internal temperatures. An intermittently run system may be run continuously or perhaps the flow water temperature can be raised slightly higher than the norm in an attempt to allow more heat to be available from the system.
Boiler control
The system has been designed to operate with a specific flow temperature, perhaps 80oC. The first control necessary for good practice is a boiler thermostat set to the required flow temperature of 80oC.
On a solid fuel boiler, the thermostat is a modulating type which increases the size of the air inlet as the boiler flow water temperature falls, and reduces it as the temperature rises. The chimney draught will influence operation, and such a thermostat is invariably calibrated in numbers rather than degrees.
With gas or oil fired boilers, an on-off control thermostat cuts off the heat source at a given water temperature, restoring combustion at a temperature a few degrees lower.
The boiler thermostat is a vital part of the system. Should it fail, the water temperature could continue to rise until danger levels are reached. For this reason, a back up ‘high-limit’ thermostat is also fitted. This would be set a few
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77
degrees higher than the control thermostat and will trip out the boiler. The high limit thermostat normally incorporates a manual re-set device.
The boiler control and high limit thermostats are fitted either in the waterway at the topmost part of the boiler or in the heating flow as it leaves the boiler.
We must be clear that the above are really safety devices, they are not intended to control the space temperature of the building being heated.
If the outside temperature falls below design conditions, any margin allowance on boiler output plus inherent margins on radiator output will help as outlined before.
More important are the days and weeks in the average winter when the temperature will be above design conditions. With the only control being the boiler thermostat, the boiler will fire to maintain the set flow water temperature, the heat emitters will give out heat and we may well end up opening a few windows, thus wasting valuable fuel and money. Alternatively, we may manually shut off several radiators but this is hardly practical.
We could also re-set the boiler control thermostat to give a lower flow temperature, but this is rather a ‘hit and miss’ affair. We may also present ourselves with an expensive problem due to boiler corrosion over a period of time, as a result of low return water temperatures.
Obviously, what is required is a further secondary automatic control over the system which will respond to temperature, thus saving fuel and maintain comfort conditions related to the ambient temperature.
We will consider this aspect next.
Hot Water Heating
TEMPERATURE CONTROL
There are many alternative methods open to us and we will look at some of these together with their relative advantages and disadvantages.
Room thermostat
This is the cheapest form of thermostatic control. The operation of the system depends upon the water being circulated by the pump around the piping system (unless we are dealing with a gravity system, of course). Two options are possible with the room thermostat. The first entails switching the pump off and on to stop water circulating; the second is to allow the pump to run continuously and switch off the boiler. The second option is sometimes preferred because it allows water to circulate through the boiler at all times, which can help to reduce thermal stress in the boiler body.
The room thermostat must be fixed in a representative position and it is often difficult to achieve this. The thermostat is only concerned with keeping the temperature constant in its immediate vicinity and must not be placed where it can be unduly influenced by extreme local conditions (for instance above a radiator). The radiator in the room containing the thermostat must never be turned down, otherwise the thermostat will not receive a representative signal and will try to keep the heating on, overheating the remainder of the space being heated. The use of some auxiliary heating, such as an open fire, will have the opposite effect.
Individual radiator control The use of thermostatic radiator valves (Fig. 45) which contain an inbuilt (or sometimes remote) temperature sensor will continuously vary the flow of hot water to individual radiators to maintain desired space temperatures. They will also compensate for local heat gains.
Fig. 45 A typical Thermostatic Radiator Valve
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Hot Water Heating
On large systems where such controls are used, some thought must be given to their affect on the hydraulic dynamic of the system. If they all close off together, pumping problems and noise can result.
Mixing valves with compensator (i.e. Variable Temperature (‘VT’) Circuits)
A three-way mixing valve is fitted in the flow line from the boiler. The mixing valve has two inlets and one outlet, which is constantly open. The position of the valve internals determines the percentage flow through each inlet. A compensating control system operates the motor on the valve. This comprises a flow water temperature sensor and an outside temperature sensor plus a control box, as depicted in Fig. 46.
Compensator Control Box or Building Management System (BMS)
Outside Sensor
Flow Sensor A C
Three Way BMixing Valve
Heating System
Alternative Pump Position Fig. 46 A Mixing Valve and Compensating Control System The control box contains a schedule which relates the desired flow temperature with the outside temperature. The schedule may, for example, be set such that flow temperature will be 80oC when the outside temperature is -1oC. As the outside temperature rises above -1oC, so the flow sensor is automatically re-set to give a slightly lower flow water temperature. To achieve this, the amount of opening in Port A is reduced, which increases the opening of Port B. This allows more of the cooler return water from the system to flow through Port B and out of the common Port C. The controller schedule is usually adjustable to cater for the different heat output characteristics of different types of heat emitter, as shown by the graphs in Fig. 47. Note that in Fig. 46, the pump is situated on the side of the mixing valve which has water flowing constantly.
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Flow
Wat
er T
empe
ratu
re o C
80
Slope
40
-1 18 Outside Temperature oC
a) Typical Schedule
Flow
Wat
er T
empe
ratu
re o C
Convectors
Floor Coils
Radiators
Outside Temperature oC b) Varying Output Characteristics
Fig. 47 Different Compensator Schedules
Diverting valves These too are three-way valves, but are piped into the circulation system such that there is only one inlet (Port A), but two outlets (Ports B & C). The amount of opening of ports B and C is varied to effect control of the piece of equipment by diverting some flow water (or even all of it) away from the plant and back through the boiler.
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Actuator (Motor)
Three Way Diverter Valve
Piece of Heating Equipment (i.e
Heater Battery)
Alternative Pump Position
A
Constant Temperature (‘CT’) circuit
C
B
Immersion Sensor
Fig. 48 Typical Diverting Valve Application Note again that the pump must be sited on that part of the circulation system that constantly has water flowing in it, regardless of the internal position of the three-way valve.
Diverting valves would normally be used to control large single pieces of equipment rather than a distribution system.
Zone control
In some buildings, the heating system can be split into separate areas, generally referred to as zones (typically north, south, east and west). Each zone. Each zone may be controlled by a separate compensator with its own outside sensor positioned in the relevant aspect. Thus the effect of solar gain on each of the zones can be taken into account.
However, a compensator cannot account for internal incidental heat gains. Sometimes an overriding internal room thermostat is added to a compensator control to cater for this. But an alternative control solution may take the form of a zone control, either with a two-way or three-way valve, with its own room sensor. We have seen that three-way valves can be used in mixing or diverting applications, as depicted in Fig. 49 and Fig. 50.
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Two Inlets, One Outlet, Constant Volume, Variable Temperature
Fig. 49 Three-Way Mixing Valve
One Inlet, Two Outlets, Variable Volume, Constant Temperature
Fig. 50 Three-Way Diverting Valve Unlike three-way valves, a two-way valve, as shown in Fig. 51, has a throttling action, and will significantly alter the volume of water flowing, and the pressure conditions of the system.
Fig. 51 Two-Way Valve
Time control
As well as controlling temperature, we may also need to control the operating time of the system. For example, if an office block is only occupied between 09.00 and 17.00, it would be wasteful to fully heat it though the night.
With a compensator control, night set back could be provided, such that the night time flow water temperatures are automatically depressed below normal temperatures during unoccupied times.
Alternatively, a simple time switch may be wired into the system to automatically start and stop the boiler plant firing at pre-determined and pre-set times.
The difficulty here is in determining the best time to start the boiler firing. Ideally it needs to be running long enough to bring the building up to the required comfort temperature just at the time of occupancy. Any time longer than this, and the building has been heated unnecessarily and fuel has been wasted.
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This has led to Optimum Start control, which, in simplistic terms, is nothing more than a sophisticated time switch, that calculates the optimum time every morning to start the heating system. It is also possible to stop the boiler firing at the best time at the end of the day, allowing the residual heat in the building to keep the occupants comfortable during this period.
Frost protection
Once time control is introduced it is necessary to protect the water filled system against the possibility of freezing and rupturing, during the ‘quiet’ hours. This is often seen to be done by means of another internal thermostat set to a typical value of 10 oC, which also protects against condensation in the building.
Building Management Systems (BMS) or Building Energy Management Systems (BEMS)
In modern buildings, be they hospitals, institutions or offices, the control of all heating and ventilation systems, fire systems, etc, is often undertaken by BMS or BEMS systems. These systems come in various types but are generally based on local controllers within each plantroom which are then networked by a central control station. The systems, via electronic controls and software, control each individual piece of plant and equipment, though in some cases they simply take alarms and provide enable/disable signals to equipment with stand-alone controls. The major benefit of these types of system is that a good centralised control system can enable optimum energy usage.
Sizing the control valve
This can become a very involved subject. Different types and makes of control valve have different flow characteristics. Manufacturers quote flow coefficient values in terms of Kv, Av and Cv (beware the potential confusion between Cvuk and Cvus values, which are based on two separate volumetric values). These coefficients are further defined in the Learning Centre Module 6.2 – Control valve capacity. The metric term Kv refers to the volume of water between 5oC and 40oC that will flow though a valve when subjected to a 1 bar pressure drop. We will often see the term Kvs; this refers to the valve capacity when the valve is fully open, and it is this figure that manufacturers quote for their valves.
For the purpose of this document, we will keep to a more simplified approach and leave the deeper subject of control valve performance to the Learning Centre Modules, especially Module 6.3.
In order to size any valve, we need to know how much water has to pass through it – the design flowrate. We also need to know the pressure conditions with which it is going to operate.
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Static pressure and differential pressure
It is important not to confuse differential pressure with static pressure.
Static pressure
The latter is the pressure inside the system due to the height of the uppermost part of the system – the water level maintained in a header tank in a feed/expansion tank system, or perhaps the pressure provided by some other means. Static pressure exists at all times whether the pump is running or not, and is equal at all points in the system which are at the same level. This means that a control valve in a horizontal pipeline is subjected to the same static pressure.
Differential pressure
The water in a heating system is continuously recirculated in a closed piping system by means of a pump. The pressure generated by the pump provides the necessary mechanical energy to circulate the water, and this energy is dissipated around the entire circuit. The energy dissipation not only takes place along the piping, in terms of pressure loss caused by friction, but also when passing through control valves, fittings, and any other item of plant. These pressure losses can be thought of in terms of differential pressure as a pressure difference exists across the pipe, fitting, or valve in the direction of flow.
In the case of control valves, we cannot size the valve until we know the differential pressure across the valve when the valve is fully open. The larger the valve, the less the differential pressure needed for the same flow; equally, the greater the flow provided by the same differential pressure. We also need to ensure that the valve actuator can provide enough force to shut the valve when required, otherwise control will suffer. In a two way valve, as depicted in Fig. 52, the closing valve will steadily increase the differential pressure to a maximum when the valve is fully shut. At this point, the pressure immediately upstream of the valve will be the pump pressure minus the friction loss in the piping between the pump and our valve. The pressure on the downstream side of the valve will depend on the friction loss in the return piping.
DifferentialPressure
Control System Force
Fig. 52 Differential pressure across a two-way valve
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Hot Water Heating
For simplicity, we can ignore the pressure in the return piping, and simply consider the pressure immediately upstream of our valve as being the highest differential pressure against which it has to close. For example, in Fig. 53, our two-way valve is used to control the flow in a branch line off a main distribution pipe.
Control Valve
Fig. 53 Differential Pressure Example across a Two-Way Valve Consider the pump develops an output pressure of 50 kPa, and the friction losses along the piping to our branch line amount to 10 kPa, then the pressure immediately upstream of our valve is simply 50 – 10 = 40 kPa. We must select a valve with an actuator that can close against 40 kPa.
We must also ensure that the valve body and internals are capable of withstanding the ‘bursting’ pressure exerted by both the differential pressure and the static pressure.
Consider our valve is part of a system that is pressurised by a header tank at a height of 5 metres above it. The static pressure on the valve will be the equivalent of a 5 m water head, which is just a little less than 50 kPa.
It can be seen that the valve has to be able to withstand a maximum pressure of 40 kPa (differential) plus 50 kPa (static) = 90 kPa in total.
As suggested earlier, we can use control valves of different sizes to deal with the relationship between differential pressure and flowrate.
If we select a large valve, the water will be pumped through it with little resistance and therefore generate little differential pressure in doing so. In contrast, we could select a very small valve and rely on a much higher differential pressure to force the same amount of water through it as the larger valve, as shown in Fig. 54.
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Flow
Pump X l/s 50 kPa
50mm Valve (Pressure drop say 2 kPa)
Pressure available (50-2) = 48 kPa
25mm Valve (Pressure drop say 40 kPa)
Pressure available (50-40) = 10 kPa
Fig. 54 Size of Valve versus Pressure Drop Superficially, we might think it better to use the smaller valve, as it will probably be cheaper, but there is a drawback. The smaller valve will require a larger proportion of the available pump pressure to ensure the same flowrate through the valve. This might cause us to fit larger piping or a larger pump to compensate. If so, any economical benefit might quickly disappear.
Correct sizing is based on some compromise between pressure drop across the valve and pressure drop across the rest of the piping system, and the relationship between these two is known as the valve authority.
Valve authority
Assume that in a particular water circuit, the available pressure immediately upstream of a control valve is 30 kPa at maximum flow (when the control valve is fully open) of which the differential pressure across the valve itself is 1 kPa. When the valve closes, the differential pressure becomes 30 kPa.
Here, the differential pressure across the valve increased by a factor of 30, from the valve being fully open to fully closed (or to fully divert or fully mix in the case of a three-way valve).
This increase in differential pressure as the valve closes can have a large effect on the ‘balance’ of the rest of the system
It is important to size the valve in a manner that this differential pressure factor is kept as low as possible. Let’s assume in the example that we have used relatively large piping and the frictional resistance is only 20 kPa. We select a smaller valve to pass the same maximum flowrate but with a 10 kPa differential pressure instead of 1 kPa as previous.
The differential pressure factor is now only 2 instead of 30, and the performance of the control valve will be improved as it now has a greater influence (or authority) over the water circuit.
An oversized control valve on a water system will have little ‘authority’ and poor control might result. An undersized control valve will either increase
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pumping costs or will result in poor circulation and heat starvation in some circuits.
In general, it can be said that the differential pressure across a wide-open water valve at full flow should be a reasonable percentage of that of the circuit in which it is installed. We can define a ‘reasonable percentage’ a little further.
Valve authority (N) is defined as the ratio of the differential pressure across the fully open valve to the pressure drop across the entire circuit, including the valve, as shown in Equation 15.
1
1 2
1
2
PN Equation 15
P Differential pressure across the fully open valve
P Pressure drop across the rest of the piping system
=
=
P P
Where :
=+
Values of N close to 0.5 give the best result, whilst lower values give poor authority. It is generally regarded that the Value of N should never be less than 0.2.
As an example, a 12 kPa pump head is available to overcome a total circuit resistance including the control valve. The differential pressure across the valve is 5 kPa, leaving a pressure drop of 7 kPa to be absorbed by the rest of the system.
From Equation 15, the valve authority will be:
1
1 2
1
2
PN Equation 15
P Pressure drop across the rest of the piping system = 7 kPa
Therefore:
5N
5 7N 0.42
=
=
=+
=
P P
Where :
P Differential pressure across the fully open valve = 5 kPa
+
=
Figures 55, 56, and 57 show the various relationships between P1 and P2 for two-way throttling applications, and three-way mixing and diverting applications. It is assumed that the mixing line resistance equals the that of the load circuit in Fig. 56, and that of the boiler circuit in Fig. 57.
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P1 P1
P2 P2
Fig. 55 Valve Authority - Fig. 56 Valve Authority - Two-Port Control Three-Way Mixing
P1
C
A B
R1P2 R2
Fig. 57 Valve Authority - Fig. 58 Three-Way Mixing Valve - Three-Way Diverting Unbalanced
The need for balancing three-way valves
Essentially, the action of a three-way (or 3-Port) valve, whether mixing or diverting, does not much influence the pressure dynamic around the whole system. This statement must be qualified, however, on the very real assumption that correct and proper balancing has been achieved. From Fig. 58, the resistance R1 in the mixing line will be somewhat less than the resistance R2 of the boiler line (which is the same size). If the mixing line resistance is not balanced with that of the boiler line, when the valve moves to a fully mixed position (Port A closed, Port C open), the pressure conditions will change as shown in Fig. 59.
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Valve to Flow Pump Curve
P1
Valve to By-Pass
P2
System Curve
Pre
ssur
e
Flow Q1 Q2 Fig. 59 Three-Way Control Action with Unbalanced Mixing Line
Fig. 59 shows us that, due to the lesser resistance in the mixing line, when the valve is fully mixing, the flow in the load circuit will be higher. We could think of the operation as having two system curves; one for the boiler circuit and one for the load circuit.
At the higher flow through the mixing circuit, the pump pressure reduces in accordance with the pump curve, and if there were many load circuits in parallel, those further from the pump could suffer from water starvation. To alleviate this problem, an artificial resistance is added to the mixing circuit via a balancing valve, as shown in Fig. 60, where R1 = R2.
R1R2
Mixing R1 = R2
R1 R2
Diverting Fig. 60 Balancing Three-Way Valve Circuits
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The need for balancing is required whether the system is mixing or diverting. If diverting, the balancing valve is placed in the diverting line. The mixing and diverting lines are sometimes referred to as ‘balance legs’.
Effects of two-port valves
A three-port valve, whether mixing or diverting, can be considered as a ‘constant volume’ valve, whereby constant pressure distribution will be maintained in the system, irrespective of the position of the valve, as long as the system is balanced. In contrast, if a two-port valve is used, then as the valve closes so flow will decrease and the differential pressure across it will increase. With the valve tight shut, dead end conditions would occur and to avoid this, a small bore ‘shunt’ pipe can be fitted across the flow and return mains, as shown in Fig. 61.
Small bore shunt pipe with valve
Fig. 61 Small Bore Shunt around a Two-Port Control Valve This merely allows water to continue circulating through the distribution system when the control valve is shut. The flow through the shunt is so small that it cannot prevent changes in pump head from occurring as the control valve closes. The effects are illustrated in Fig. 62.
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Valve in a partly closed position
Valve Pressure
Valve Pressure Drop
Pipe Pressure Drop
System Pipe Pressure Drop
Valve Fully Open
Drop
Increased Pressure
Design Pressure
Flow
Pre
ssur
e
Fig. 62 Effect of a Two-Port Valve on Pump Pressure This change in pump pressure may be deliberately required, such that the pressure changes are used to trip in and out, two or more different sized circulating pumps in sympathy with changes in heat requirements. One other alternative is to use a pump with a flat characteristic curve, so that the reduction in flowrate as the control valve throttles will only result in a small change in pressure.
Reduced Flow Design Flow
Multi-zone circuits with three-way control Fig. 63 shows a series of zones each complete with a mixing valve.
Fig. 63 Three Zone Circuit with Common Pump
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There is only one pump serving the whole network. This arrangement cannot work, however. When any one heating load requires water at a temperature below that of the boiler flow temperature, the relevant control valve partly closes the supply port from the boiler, and opens the port from the mixing leg. The pump pressure then acts to push the boiler water into the return line, bypassing the heating load, instead of mixing return water with the flow to the heating load Each mixed water circuit should have its own pump with the primary circuit merely supplying hot water to small distribution headers, A, B, and C in Fig.64.
Fig. 64 Three Separate Zone Circuits with Separate Pumps The distribution headers should be larger than the distribution flow and return pipes, keeping the pressure drop across the header as little as possible. Thus, the flow in the primary circuit will not affect the flow in the secondary circuit and vice versa.
Fig. 64 shows a correct set-up. Balancing valves on the returns from the distribution headers are installed to prevent ‘short-circuiting’ to equalise the pressure drops along the primary circuit, as discussed in a previous section.
Automatic balancing
We have discussed the various aspects of the importance and procedure for balancing systems. We have also seen how the pressures in the circulation system can vary especially with variable volume systems. Balancing such a system can be difficult.
Circuit Pump or Zone Pump Primary Pump
Heat Load
Regulating Valve
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If it is set up for one condition via manually set balancing valves, as soon as controls begin to work, the dynamic of the systems alters, and pressure conditions and balancing settings will change also. With a manually balanced installation, it is obviously not practical to continuously alter balancing settings to match the changing conditions. However, automatic balancing valves can this work with relative ease; they exist in two basic forms. The first of these is an automatic flow control valve, sometimes referred to as a flow limiter, as shown in Fig. 65. An orifice plate in the return line enables a pressure difference to be measured. The greater the flow through the orifice plate, the greater the pressure difference across it. The upstream and downstream pressures act on different sides of an actuator diaphragm, forcing the diaphragm to move relative to the flow. So, if the flow increases, the valve is configured such that the valve closes, thereby maintaining a constant flow through the circuit.
Fig. 65 Flow Limiter The second type is shown in Fig. 66, and involves the same type of valve and actuator, but is applied slightly differently to act as a differential pressure control across the heating load, rather than a separate orifice plate. In the same manner as the flow limiter, as the flow increases, the pressure drop increases across the heating load and this change is sensed by the control valve diaphragm to close the valve position and maintain a constant flow through the circuit.
Fig. 66 Differential Pressure Control
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HIGH TEMPERATURE HOT WATER SYSTEMS We will not deal with this subject in depth since a considerable degree of design expertise is required, but we will outline the principle considerations. Fig. 67 shows the relationship of temperature with pressure of boiling water. The graph show, that for higher pressures, it is necessary to heat the water to higher temperatures before steam will be produced. This basic fact is used to advantage in pressurising systems to obtain water at temperatures above 100oC.
Tem
pera
ture
o C
Pressure barg
Fig. 67 Pressure/Temperature Relationship of Boiling Water An artificial static head is applied on the system, and early systems used a steam space or cushion in the boiler to create pressure. However, these days, it is more usual to employ an inert gas cushion, such as nitrogen, and packaged pressurisation units are now available in standard ranges.
So, high temperature hot water systems are those whereby the system is run at such a pressure that the flow water is maintained at a temperature in excess of its atmospheric boiling point by the pressure of steam or gas in contact with it, and circulated in a closed system by means of centrifugal pumps.
It is evident, when using pressurised systems, that a danger exists because the water could easily ‘flash’ into steam, should the system pressure fall in any way. As a precaution against the formation of steam, it is essential to apply a temperature margin, known as the ’anti-flash’ margin, generally about 15oC.
The definition of anti-flash margin is: ‘The difference between the temperature of the water in circulation at any point in an installation and the temperature at which the water would boil at the pressure sustained at that point’.
The anti-flash margin should be applied to the most vulnerable point in the system. The determination of this point can be fairly complicated and is affected by the position of the pump in the circuit.
It is standard practice to site the pump in the flow, and the connection from the pressurisation unit is taken as the datum level, or neutral point. Whatever the pressure in the boiler, the additional head created by the pump is added so that the flow main is at a pressure well above that in the boiler.
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So the system should be operated at a pressure corresponding to that necessary to prevent flashing at the desired flow water temperature, plus the anti-flash margin.
Let us assume that the water flow temperature is 150oC. The saturation pressure at this temperature is about 374 kPa gauge. Adding an anti-flash margin of 15oC gives a temperature of 165oC, and the resulting saturation pressure at this temperature is about 600 kPa. So the system would be pressurised to 600 kPa but controlled to operate at a flow water temperature of 150oC.
The design process includes producing system pressure diagrams for both the pump and static pressures and then checking the safety margin (the anti-flash margin) at the most vulnerable point in the system, which is generally the highest, should the pump stop for any reason.
Of course, the simplest method of pressurising a system is by means of a high level header tank. This application is limited though as a head of one metre is required to pressurise to about 10 kPa. In the example above, the static head of 600 kPa would require a header tank to be just over 61 m above the circulation system. Few buildings are able to accommodate this sort of head.
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DOMESTIC HOT WATER SUPPLIES
No information on hot water systems would be complete without some reference to hot water supply, commonly referred to under the general heading of Domestic Hot Water Services (DHW or DHWS).
The types of system can be divided into local and central systems. Local systems include electric water heaters or instantaneous gas fired water heaters.
DHW systems
Fig. 68 illustrates a DHW system in which any type of boiler may be used to provide the primary heat source.
Fig. 68 Un-vented DHW System
The hot water storage vessel is called a storage cylinder or storage calorifier, the former tending to be used in domestic type applications, and the latter in commercial and industrial applications.
The definition of cylinder usually refers to a copper vessel in which a copper coil is permanently fixed, whereas a calorifier tends to have a removable coil, which is handy for maintenance purposes.
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Fig. 69 A Storage Cylinder and Storage Calorifier
Calorifier - removable coil Cylinder – fixed coil
HWS storage and boiler power
Unless an instantaneous water heater is being used, such as an EasiHeat unit, it is necessary to store a certain amount of hot water for use in sinks, baths, and showers etc. A heat up time for the volume of stored water must be determined and the amount of heat needed to achieve this has to be included in the boiler power requirements.
Both the amount of hot water storage and the required boiler power are generally based on the use of the building, and the number of occupants as indicated by Table 15.
Table 14 Typical HWS Storage and Boiler Power Ratings Building Storage at 65oC
(litre/person) Boiler Power to 65oC
(kW/person)
Boarding School 25 0.7
Day School 5 0.1
Factories 5 1.2
Hotels 45 First Class
Average 35
1.2 0.9
Offices 5 0.1
For instance, for an office with fifty occupants, HWS storage provided should be 250 litres and the required boiler power would be 5 kW.
For the average single-family dwelling, there should be a storage vessel of not less than 160 litres, and a four-hour heat up period is generally acceptable.
The cold water header tank volume should not be less than the HWS vessel.
Secondary piping
The piping connecting the storage vessel to the draw-off points, such as basin taps etc, is termed the secondary piping. In smaller schemes, the secondary piping may not be arranged as a circulating system (i.e. domestic properties).
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It is referred to the secondary flow, and is a ‘dead-leg’, a pipe in which water lies stagnant when the tap is shut off.
In larger systems, the secondary piping will be arranged as a circulation system. There will be a secondary flow ring main, from which water is drawn by the taps, and a secondary return connecting the flow pipe back to the storage vessel. From the secondary circulation piping, short legs will supply single draw-off points or groups of draw-off points.
The purpose of the circulation is to reduce the quantity of cold water that must be drawn off (and therefore wasted) before hot water flows from the tap. With a circulation system, the water to be drawn off the ring main will already be at the required temperature when the tap is opened. The disadvantage of the ring main system is that it will usually cost more to install and, because hot water is continuously in circulation, heat losses will be higher.
The first step in designing and sizing secondary HWS piping is to determine the maximum flowrate likely to occur at any one time. In order to estimate this figure, information is required on the number, position and type of draw-off point, e.g. bath, basin, shower etc since this will have a bearing on the amount of water drawn off at any one time and at different times of the day. An assumption, based on practical research and experience, must also be made as to the number of draw-off points that are to be opened simultaneously. It is highly improbable that all taps in a building will be discharging hot water at the same time.
Calculation of hot water demand
There are various methods of calculating the demand of hot water for any application. Historically, the method laid down by the Chartered Institute of Building Services Engineers (CIBSE) used Demand Units or Simultaneous Demand methods, however, most building and environmental services engineers are now adopting the BS6700 Loading Units approach. We will therefore only consider this latter method here. For further information on the CIBSE methods, please refer to the relevant CIBSE guide.
Fig. 70 depicts 20 wash hand basins installed in various toilet facilities throughout a multi-storey office block. This will tend to have even, but small, continuous demand throughout the working day.
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Fig. 70 Multi-Storey Office Block with 20 Basins
BS6700 Loading Units
In most buildings appliances are rarely in simultaneous use, therefore for reasons of economy, it is usual to provide for a demand less than the total demand of all appliances being in use at the same time. The simultaneous demand can be determined from data derived by observation and experience of similar installations, or by the application of probability theory.
BS6700 Loading Units is a system of determination based on probability theory, which take into consideration the flow rate required at the appliance, the length of time in use, and the frequency of use. The number of each type of appliance, fed by the length of pipe being considered, should be multiplied by the loading units, as given in Table 15, and the total Loading Units derived for the pipe. Using Figure 71 the total number of Loading Units can be converted into the total simultaneous demand for the pipe in litres per second.
Table 15 is based on normal domestic usage and customary (or statutory) provision of appliances. It is not applicable where usage is intensive, for example, in theatres and conference halls; in such cases, it is necessary to establish the pattern of usage and appropriate peak flow demand for the particular case.
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Table 15 Loading Units – Hot or Cold Supply Type of Appliance Loading Unit (LU)
WC Flushing Cistern 2
Wash Basin ½” – DN15 1.5 to 3
Bath Tap ¾” – DN20 10
Bath Tap 1” – DN25 22
Shower 3
Sink Tap ½” – DN15 3
Sink Tap ¾” – DN20 5
Domestic Clothes or Dishwashing Machines ½” – DN15 3
NOTE 1: WC cisterns with either single or dual flush control have the same LU. NOTE 2: The wash basin LU is for use where pillar taps are installed. The larger LU is applicable to situations such as schools and those offices where there is a peak period of use. Where spray taps are installed, an equivalent continuous demand of 0.04 l/s should be assumed. NOTE 3: Urinal cistern demand is very low, and is normally disregarded. NOTE 4: Outlet fittings for industrial purposes or requiring high peak demands should be taken into account by adding 100% of their flow rate to the simultaneous demand for other appliances obtained by using LUs.
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Fig. 71 Conversion of Loading Units to Design Flow Rate Calculations for Fig. 70
From Table 15, a figure between 1.5 to 3.0 should be used for wash hand basins, so in our case we will use 2.5 Loading Units (based on Judgement of the application).
Therefore, 20 basins x 2.5 Loading Units per basin = 50 Loading Units.
From Fig. 71, converting 50 Loading Units, results in a design flow rate of approximately 0.8 l/s. Working within normal pipe sizing criteria this will equate to a pipe size of 35mm Cu leaving the calorifier.
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A system should be designed so that the design flow rates given in Table 16 are available at each outlet and any group of outlets where the total demand does not exceed 0.3 l/s, when only that outlet or group of outlets are open. When simultaneous discharge occurs the rate of flow of water at any outlet in use should not be less than the minimum rate given in Table 16.
Simultaneous use of appliances may reduce flow rates, possibly below design values. It is important therefore that the whole system should be designed so that flow rates are not reduced to such an extent as to adversely affect the satisfactory functioning of the system. In particular, where the reduction in flow could affect the temperature of water delivered to showers, measures should be taken to protect the user against excessive water temperatures.
Table 16 Design Flow Rates Rate of Flow (l/s)
Outlet Fitting or Appliance Design Rate Minimum Rate
WC Cistern (to fill in 2 minutes) 0.13 0.05
WC Flushing Trough (per WC served) 0.15 0.10
(see NOTE 2)
Urinal Cistern (each position served) 0.004 0.002
Wash Basin 0.15 0.10
Hand Basin (pillar taps) 0.10 0.07
Hand Basin (spray or spray mixer taps) 0.05 0.03
Bidet 0.20 0.10
Bath (¾”) 0.30 0.20
Bath (1”) 0.60 0.40
Shower Head (see NOTE 3) 0.20 0.10
Kitchen Sink (½”) 0.20 0.10
Kitchen Sink (¾”) 0.30 0.20
Kitchen Sink (1”) 0.60 0.40
Washing Machine 0.20 0.15
Dishwashing Machine (see NOTE 1) 0.15 0.10
Pressure Flushing Valves for WCs or Urinals 1.5 1.2
Urinal Flushing Cistern 0.3 0.15
NOTE 1: The manufacturer should be consulted for required flow rates to washing and dishwashing machines for other than single dwellings. NOTE 2: WC flushing troughs are recommended where anticipated use of WCs is more frequent than once per minute. NOTE 3: The rate of flow required to shower heads will depend on type fitted and the advice of the shower manufacturer should be sought.
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Secondary Circulation Pumps The purpose of the secondary circulation pump is two fold;
• to prevent the wastage of water by ensuring that, within a reasonable time, hot water is available at the outlet,
• to ensure that the return water temperature is maintained at a certain temperature (see section on Legionnaires disease).
Thus, the pump must circulate sufficient water to cater for the heat losses from the pipework system. This would normally be based on flow temperatures between 60 and 65oC with a minimum return temperature of 50oC. Thus, to calculate the pump duty, the following formula can be used;
( )21186.4 TTQm
−×=
Where; m = Required flowrate (l/s) Q = Heat loss from system/pipework (kW) T1 = Flow temperature (oC) T2 = Return temperature (oC)
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LEGIONELLOSIS
Background Legionellosis is the collective term for all illnesses caused by Legionella bacteria. There are approximately 48 species of Legionella bacteria with 16 Serogroups of Legionella pneumophilla (L.pneumophilla), only Serogroup1 causes Legionnaires disease whilst others cause milder illnesses such as Pontiac and Lochgoilhead fevers which are more flu like in their symptoms. Legionnaires disease was first recognised in July 1976 following an outbreak at the Bicentennial American Legion Convention, hence the name, in Philadelphia when there were 182 cases of pneumonia reported with a higher than expected number of deaths, 29 in all. The principal route of infection is through the inhalation of aerosols, fine water droplets containing the bacteria. Under ideal conditions Legionella can survive in aerosols for 2-3 hours allowing the bacteria to be transported through the air to potential victims. Once inhaled the bacteria enter the alveoli deep within the lungs where it will multiply and infect the host. Sprays, water impacting surfaces and bubbles bursting can create aerosols. There is no evidence of person-to-person transmission. Factors increasing susceptibility of contracting the disease include; • People over 50 years old, children rarely infected • Sex – males 3 times more susceptible • Existing respiratory disease • Smokers • Immuno-suppressed patients such as cancer treatment Legionella is a ubiquitous organism freely existing in nature in rivers, stream, puddles and soil. It only presents a health risk once it infects a water system, most commonly through the mains supply, which can generate aerosols. For the bacteria to proliferate within an infected system it requires the following: • Temperature of 20 – 45oC, optimum being 37oC body temperature • Presence of metal ions, notably iron from corrosion of system • Presence of amino acids, from other microbes and organic fouling • Reduced oxygen and elevated carbon dioxide • In essence a dirty system Based on these requirements there are some basic precautions for Domestic Water Services to reduce the risk from Legionella;
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Do • Store and distribute cold water below 20oC • Store hot water above 60C and distribute at >50oC • Regularly inspect, clean and disinfect storage tanks • Regularly clean and pasteurise calorifiers • Disinfect new and refurbished systems Don’t • Allow cold water to rise above 20oC • Allow hot water below 50oC • Allow dead-legs in system • Allow water to stagnate in unused parts of the system
Regulation and Enforcement Legionnaires disease is a reportable illness under the ‘Reporting of Injuries, Disease and Dangerous Occurrences’ regulations 1995. Anyone operating water systems that are at risk must comply with the Approved Code of Practice (ACoP), L8 issued by the Health & Safety Commission put into force in January 2001. The document gives practical guidance on, and has legal status under, the Health & Safety at Work etc Act 1974 and COSHH regulations with regard to risk from exposure to Legionella bacteria. Failure to comply with L8 has resulted in prosecutions. The ACoP applies to any system where water is used or stored and where there is a means of creating and transmitting water droplets that may be inhaled. Such systems include cooling towers, domestic hot and cold systems, spa baths and pools. Legal duties under L8 include; identifying and assessing the sources of risk, preparing a written scheme for prevention and controlling the risk, implementing and managing precautions, keeping records of precautions carried out and appointing a person to be managerially responsible. A risk assessment should consider the source of water supply, possible contamination, the entire system and its mode of operation including possible changes to normal use, schematic drawings and asset register, previous records and results and people using the system or their locality to the system. Control measures to prevent proliferation of Legionella and reduce exposure to aerosols should include; control release of water sprays, control water temperature, avoid stagnation, avoid materials of construction that encourage growth and ensure correct operation, regular maintenance and cleaning of the system.
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Records must be kept regarding the management and maintenance of the system, these must include a copy of the Risk Assessment, which must be reviewed every 2 years, and details of person responsible for conducting it, a copy of the Written Scheme and details of person responsible for managing and implementing it, and results of monitoring, testing and checks conducted on the system. These records must be kept for a minimum of 5 years. Responsibilities of suppliers under L8, including Spirax Sarco; • Products and services to be effective and safe • Provide information on correct and safe use of products • Make known limitations in expertise and scope of products and services • Report deficiencies, limitations or matters of concern relating to your
systems • Ensure their staff are competent
Maintenance Precautions and Control Measures In keeping with the temperatures required for Legionella to thrive within a Domestic Water System, the principal means of controlling Legionella is by maintaining strict temperature regimes throughout. • Cold water to reach cold taps at a maximum of 20oC within 2 minutes • Hot water to leave calorifiers at minimum 60oC and return at minimum
50oC • Hot water to reach hot taps at minimum 50oC within 1 minute • Consider risk of scalding Domestic Hot Water calorifiers present an increased risk from Legionella if not correctly maintained - our EasiHeats greatly reduce this risk and simplify ongoing compliance to L8. Calorifiers need to be in good condition and clean, to reduce corrosion debris and sludge fouling which are growth requirements of Legionella. They need to have access for regular inspection and cleaning under L8, water must be circulated through out the calorifier to ensure even temperature and inlet outlet temperature gauges are required for recording flow and return in accordance with L8. Particular care must be given to standby units, these need to be pasteurised at 60oC for a minimum 1 hour before returning to service. Records must be kept in a logbook detailing manufactures instructions, schematic drawings showing all main valves and mode of operation, total system volume, information on any water treatment regime such as chlorine dioxide, all results of temperature monitoring, control limits and corrective actions and all cleaning and disinfection procedures. Domestic systems should be cleaned and disinfected if routine inspection shows it to be necessary, if the system or part of it has been substantially
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altered or entered for maintenance purposes, or following a suspected outbreak of Legionellosis.
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SUMMARY
As suggested at the very beginning, this document is not intended to convert anyone into a fully fledged heating system designer or building services practitioner; indeed, any single chapter of this document could be expanded into a full text book in its own right.
Rather, we hope that we have dealt with the basics of hot water heating systems in a way that will increase our general understanding of the design and dynamics of such systems.
The basic information in this document should arm us with enough confidence to deal with simple heating systems, and provide a sturdy foundation to a higher level.
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