70 40 - sav-systems.com · • BSRIA Commissioning Water Systems guide BG2/2010 • BSRIA Guide...

www.sav-systems.com Rev: 2.0 10/2014

‘Delta T’ Design Guide

Low Carbon System Design - a whole system approach

“70ºC fl ow / 40ºC return”

70 40

References

• Danfoss Technical Paper: Results and experiences - Lystrup 12/2013

• CIBSE Knowledge Series Guide KS7: Variable flow pipework systems 2004

• CIBSE Commissioning Code W: Water Distribution Systems 2010

• CIBSE AM12 2013

• BSRIA Commissioning Water Systems guide BG2/2010

• BSRIA Guide BG12/2011: Energy Efficient Pumping Systems 2011

• SAV FlatStation Design Guide 2014

• SAV LoadTracker CHP Design Guide 2014

• GLA’s District Heating Manual for London 2013

This document is based on the best knowledge available at the time of publication. However no responsibility of any kind for any injury, death, loss, damage or delay however caused resulting from the use of these recommendations can be accepted by SAV Systems, the authors or others involved in its publication.

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‘Delta T’ - DESIGN GUIDE70 40

SYMBOLS

ISOLATING VALVE

4 PORT CONTROL VALVE

PRESSURE TEST POINT

NON-RETURN VALVE

DRAIN OFF COCK

CONSTANT FLOW REGULATOR

PRESSURE GAUGE

TEMPERATURE GAUGE

FLEXIBLE HOSE

STRAINER

FLEXIBLE COUPLING

THERMOSTATIC RADIATOR VALVE

LOCKSHIELD VALVE

DIFFERENTIAL PRESSURECONTROL VALVE

AUTOMATIC AIR VENT

FIXED ORIFICE FLOW MEASUREMENTDEVICE (ORIFICE PLATE)

DOUBLE REGULATING VALVE

FIXED ORIFICE DOUBLE REGULATINGVALVE (COMMISSIONING SET)



PUMP

PRESSURE INDEPENDENTCONTROL VALVE

SAFETY RELIEF VALVE

BLANKED FLANGE PIPE END

Guide’s technical author, Chris Parsloe of Parsloe Consulting Ltd

Chris is the author of numerous industry guides on variable flow heating

and chilled water systems including:

• CIBSE Knowledge Series Guide KS7: Variable flow pipework systems 2004

• CIBSE Commissioning Code W: Water Distribution Systems 2010

• BSRIA Guide BG12/2011: Energy Efficient Pumping Systems 2011

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1. INTRODUCTION

In order to realise the predicted energy saving benefits of commercial heating systems and heating networks, it is essential that pipework distribution systems be designed with proper regard to system operating temperatures.

This design guide makes frequent reference to ‘delta T’, which is the difference in temperature between the flow and return water in a piped heating distribution system. The selection and maintenance of system flow and return temperatures requires close attention if the energy and cost saving potential of a system are to be realised.

In those cases where the anticipated energy savings of heating distribution systems do not materialise, these failings can frequently be traced to the inability to maintain ‘delta T’.

There are three key objectives of successful ‘delta T’ designs:

• The operating ‘delta T’ across heat sources should be as close as possible to their optimal design values. This will enable them to operate at close to their peak efficiencies.

• The heat dissipation from, and hence ‘delta T’ across secondary distribution circuits should be maximised. This will give rise to lower system flow rates, smaller pipes and smaller pumps that consume less energy. Furthermore, buffer vessels or thermal stores can be reduced in size relative to the size of the system.

• Secondary flow and return water temperatures must be kept at as low as possible. This will help to minimise pipe distribution heat losses.

Some designers take the view that to influence system operating temperatures is impossible and that other considerations should take priority. This guide challenges this view by explaining how good ‘delta T’ design can be undertaken in a way that delivers significant energy and cost saving advantages.

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2. PRIMARY CIRCUIT DESIGN

Proper design of the primary circuit is essential. Most low carbon heat sources have specific flow and return water temperature requirements that enable them to perform at their optimal efficiencies. The aim should be to ensure that these conditions are maintained for the majority of the operating period.

Table 1 shows typical recommended operating temperatures for different types of low carbon heat source.

Heat source Reasons

Flow (°C) Return (°C)

Gas condensing boiler

55 30 Condensing of flue gases is only possible at temperatures of less than 55ºC and fullcondensation only occurs at less than 35ºC.

Biomass boiler 80 60 High grade heat is available from biomass boilers but return temperatures must be maintained at greater than 60ºC to avoid condensation of flue gases which might lead

to corrosion within the boiler.

CHP 80-90 70 High grade heat is available from CHP units but low flows and excessive delta T values may cause uneven cooling of the CHP unit.

Heat pump 40 35 Heat pumps produce low grade heat. The efficiency and output of heat pumps is significantly reduced at higher flow temperatures.

Optimal primary circuit ‘delta T’

NB: These values apply to localised primary flows only. Secondary circuit design should be with the lowest possible return temperatures and the largest possible ‘delta T’ values for the reasons explained under the Importance of ‘delta T’ (see page 12).

Table 1: Typical primary circuit operating temperatures

It is likely that different types of heat source will be used in the same system (such as gas boilers providing back-up for a biomass boiler or CHP unit). Where this is the case the lead heat source should be the one that has the lowest carbon emissions, as recommended by Part L of the Building Regulations. It then follows that the primary circuit flow temperature should be dictated by the lead heat source e.g. if water at 80°C is available from a CHP unit, then this should be the design flow temperature for the primary circuit. However, for reasons that will become clear, secondary circuit temperatures at terminals should be as low as possible.

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Figure 1 shows a primary / secondary circuit layout that enables this objective to be achieved. The following notes apply to each of the labelled features.

PRIMARY INJECTIONCIRCUITCIRCUIT

PRESSURISATIONUNIT

BUFFER/THERMAL

WEATHER COMPENSATION

DELTA ‘T’CONTROLLER

STOREAUXILIARY ANDBACK-UP HEAT

SOURCES

AUXILIARY HEATSOURCELEAD HEAT

SOURCE

A

A

AE

GB

B

B

F

D

H

C

Figure 1: Primary circuit schematic showing low carbon heat source options

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A. All heat sources connected to the primary circuit should be fitted with a circulating pump and mixing valve arrangement. This is for one or both of the following purposes

a. To ensure that the flow temperatures from each heat source are held constant and at the same set point value - this being dictated by what best suits the lead heat source. The maintenance of a constant combined flow temperature from heat sources is important for proper functioning of buffer vessels and thermal stores. If the flow temperature from a heat source is ever less than the set value (as might be the case during system start-up conditions) flow water can be mixed in with return water until the required flow temperature is achieved.

b. To ensure that the return temperature does not drop below a minimum value as specified by the heat source manufacturer. This is particularly relevant in the case of biomass boilers that require a minimum return temperature of 60°C to prevent internal corrosion. As for (a) above, if the return water temperature is less than the required value, flow water can be mixed in with return water until the required return temperature is achieved.

B. Each of the pumps serving the various heat sources should be sized such that it is capable of delivering the design flow rate through the heat source it serves against the design pressure loss in the entire circuit between the heat source and the thermal store. These pumps should be variable speed but, during commissioning, they should each be set at a fixed speed sufficient to achieve the required design flow rate through the heat source when circulating through the thermal store. The duties of pumps supplied as integral to the heat source should be checked to ensure that they have sufficient capacity to deliver flow to the thermal store.

C. A “de-coupler” should be included that includes a buffer tank or thermal store. The direction of flow through the de-coupler and storage vessel may be in either direction, depending on the system operating condition.

D. The secondary pumps should be variable speed pumps sized to deliver the anticipated maximum design flow rate against the secondary circuit design pressure loss. The flow temperatures for each secondary circuit should be at as low a value as possible and each circuit should have the means to vary the flow temperature to suit the particular operating conditions. Constant temperature secondary circuits should never be used since they risk delivering high temperature return water back into the primary circuit.

E. Where operating temperatures are compatible, the lead heat source can be supplemented by a heat source located in the main system return that pre-heats the water before it reaches the lead heat source. For example, this opportunity exists in a system where the lead heat source is a CHP unit that has been selected to heat water from 70°C to 80°C whilst secondary circuits are designed to deliver a return temperature of 40°C. In this case a heat pump might be used to pre-heat the returning water from 40°C to 45°C or, alternatively a gas condensing boiler might be used to pre-heat the water from 40°C to 55°C. It can be seen that this enables the auxiliary heat source to operate at temperatures that achieve optimal efficiency. The heat source must include its own pump capable of injecting heated water back into the return pipe. This means that the heat output available is limited by the design flow rate in the primary circuit - since the flow rate through the heat source cannot exceed the design flow rate in the primary circuit.

F. Additional auxiliary or back-up heat sources can be connected in parallel with the lead heat source. An auxiliary boiler will supplement the heat output of the lead heat source whereas a back-up boiler will only run when other heat sources are not available. Since each of these heat sources must achieve the same flow temperature as the lead heat source, it may be necessary to diverge from optimal operating temperatures. For example a condensing boiler operating as an auxiliary or back-up to a CHP unit would have to operate at a flow temperature that suited the CHP unit (typically 80°C) thereby reducing the potential to operate in condensing mode. For this reason, auxiliary or back-up heat sources in these locations should be the last to operate under sequencing controls.

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G. An injection pump is required to push water from the primary circuit into the header from which secondary circuits are attached. This will ensure that the secondary circuits cannot affect flows in the primary circuit. A two port control valve ensures that the flow of water into the header is sufficient to maintain the required main secondary flow temperature. If necessary, this temperature can be further mixed to lower temperatures at individual secondary circuits. An additional sensor in the low loss header can be used to detect any rise in temperature suggesting a reversal of flow through the header. The two port valve should then throttle to reduce flow if this should happen.

H. Buffer stores and thermal stores are generally of benefit to all heating systems and are a critical aspect of system design. The following subsection explains their role in detail:

Buffer vessels / thermal stores

Buffer vessels / thermal stores are located in the de-coupler of the primary circuit. Although a single vessel often serves a shared function, the separate roles of buffer vessels and thermal stores can be considered as follows:

• Buffer vessels. A buffer vessel enables the heat source to run at its optimal return temperature for as long as required. For example, if there were no buffer vessel in Figure 1, a proportion of the heated water would inevitably short circuit through the de-coupler back to the heat source thereby raising the temperature of the return water. For condensing boilers, this might force them into a non-condensing condition reducing their efficiency. By installing a buffer vessel, each heat source can operate for a fixed period, as determined by the size of the buffer vessel, without the risk of elevated return temperatures. In the case of biomass boilers this is particularly important since elevated return temperatures may result in over-heating of the water inside the boiler. Furthermore, biomass boilers continue to generate heat for a significant period after being switched off. This residual heat needs to be accommodated by the system, so an appropriately sized buffer vessel is essential.

• Thermal stores. A buffer vessel that is designed to store heated water generated during periods of relatively low heating demand, in preparation for an anticipated high demand, is referred to as a “thermal store”. For CHP units, it may be the case that the unit runs during periods when there is a demand for electricity but little demand for heat. An appropriately sized thermal store enables the CHP unit to run during these periods, creating a store of heated water which can then be used later on when a demand for heating arises. Alternatively, a thermal store may be required in order to deal with high short term heat loads in the system. This might typically be the case in a system serving FlatStation heat interface units (HIUs) that provide instantaneous heating of hot water. These units cause high short term energy demands during periods of peak hot water draw-off. Rather than sizing the main heat sources to cope with these short term loads, a thermal store can be used to create a store of heated water capable of dealing with a high short term demand. This benefits the sizing of the heat sources.

Buffer vessels and thermal stores can be used to control the on / off sequencing of each heat source.

Figure 2 shows a typical buffer vessel or thermal store, where the position of the horizontal separation layer between flow and return water is used for on / off sequencing. Sensors located in the side of the vessel determine the movement of the separation layer enabling the various heat sources to be sequenced on and off.

For a buffer vessel, the distance between the sensors will be dictated by the required run time of each heat source. For a thermal store, the volume of the vessel may be increased to enable the lead heat source to operate for as long as possible.

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If temp < set-point, auxiliary heat source ON (location E)If temp = set-point, auxiliary heat source location F OFF

If temp < set point, lead heat source ONIf temp = set-point, auxiliary heat source location E OFF

If temp = set point, lead heat source OFF

(red – separation layer rising = thermal store emptying)

(blue – separation layer falling = thermal store filling)

Locations E & F shown in Figure 1

If temp < set-point, auxiliary heat source ON (location F)

Volume available forresidual heat if required.

BUFFER/THERMALSTORE

Figure 2: Buffer vessel / thermal store used for heat source sequencing

In all circumstances, the key to satisfactory control is to ensure that there is good stratification of temperature within the vessel. This means ensuring that:

• The pipes entering and leaving the vessel are from the side (rather than top and bottom) in order to minimise any turbulence that might disturb temperature stratification. Furthermore, the water should enter and exit the vessel through diffusers, i.e. at reduced velocity, and be directed upwards at the top connection and downwards at the bottom connection.

• The temperature of the water entering the top of the vessel must be at a consistent fixed value. This can be achieved by including mixing arrangements on each of the main heat sources as explained in the preceding section.

• The temperature of the water entering the bottom of the vessel (when the vessel is discharging) must be maintained at a value that is at least 25°C lower than the heated water at the top of the vessel. This ensures a good separation layer between flow and return water enabling accurate measurement of the heated water level in the vessel.

The approach indicated in Figure 2 is simple to implement and provides an effective method of sequencing heat sources. Figures 3, 4, 5 & 6 demonstrate how EC Power LoadTracker CHP controls various heat sources based on thermal store separation layers.

EC Power LoadTracker CHP thermal stores take this concept further by incorporating a controller that analyses how the temperatures in the store have changed over the preceding seven days. This information is then used to calculate the optimum start time for the CHP unit thereby prolonging CHP run times.

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EC Power LoadTracker under varying load conditions

Figure 4: Medium Load (40%) - Heat Source 1 & 2 Only

Figure 3: Low Load (20%) - Heat Source 1 Only

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Figure 6: High Load (100%) - Heat Source 1, 2, 3 & 4

Figure 5: Medium Load (70%) - Heat Source 1, 2, & 3 Only

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Importance of ‘delta T’

The selection of secondary flow and return temperatures (which produce ‘delta T’) have a major impact on the installed cost and energy efficiency of a system such as that illustrated in Figures 3-6.

The implications are best explained by comparing two possible alternatives. These are as illustrated in Figures 7 and 8. For each pair of flow and return pipes the flow rate can be estimated using the equation

qm = P / (c

p DT)

where,

qm = mass flow rate (kg/s) (≈ volume flow rate in litres / second)

P = Heat load (kW)

cp = specific heat capacity of water (= 4.18kJ / kgK)

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PRIMARYCIRCUIT

BUFFER/THERMALSTORE

BACK-UP BOILER

CHP UNIT 100kW

AUXILIARY BOILER 100kW

80°C

2.4 litres/sec

2.4 litres/sec

60°C1.2litres/sec

60°C1.2litres/sec

80°C

60°C

2880 litres

60°C

60°C

SECONDARY PUMPS

80°C1.2 litres/sec

100kW

80°C1.2 litres/sec

100kW

Figure 7: System operating at ‘delta T’ of 20 (=80 / 60) in primary and secondary circuits

Figure 7 shows a primary circuit into which a CHP unit and an auxiliary gas condensing boiler, each rated at 100kW, are supplying water at 80°C. A single constant temperature secondary circuit distributes the water at 80°C to terminal units that are sized to generate a 20°C ‘delta T’ giving a return temperature of 60°C. A buffer vessel is included to store water at the 80°C flow temperature.

For this system (ignoring the intervention of heat source sequencing controls):

• The peak flow rate in the primary circuit is 2.4 l/s - i.e. 200 / (4.18 x (80-60))

• The peak flow rate in the combined secondary circuits is also 2.4 l/s (since load and ‘delta T’ are the same as for the primary)

• A buffer vessel sized to provide a notional 20 minutes of flow at 80°C would need to have a volume of 2880 litres i.e. 2.4 l/s for 20 minutes.

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PRIMARY

INJECTIONCIRCUIT

CIRCUIT

PRESSURISATIONUNIT

BUFFER/THERMALSTORE

BACK-UP BOILER

CHP UNIT 100kW

AUXILIARY BOILER 100kW

80°C

1.2 litres/sec

1440 litres

1.2 litres/sec

40°C0.6 litres/sec

70°C

40°C

0.8 litres/sec100kW

40°C

80°C

40°C0.6 litres/sec

40°C

SECONDARY PUMPS

AUXILIARY HEATSOURCE

70°C0.8 litres/sec

100kW


DELTA ‘T’CONTROLLER

Figure 8: System operating ‘delta T’ of 20 (=80 / 60) primary and 30 (=70 / 40) secondary circuits

Figure 8 shows the same primary arrangement but with a variable temperature secondary circuit where the flow temperature is 70°C and terminal units are sized to generate a 30°C ‘delta T’ giving a return temperature of 40°C.

For this system:

• The peak flow rate in the primary circuit is 1.2 l/s - i.e. 200 / (4.18 x (80-40))

• The peak flow rate in the combined secondary circuits is 1.6 l/s - i.e. 200 / (4.18 x (70-40))

• A buffer vessel sized to provide a notional 20 minutes of flow at 80°C would need to have a volume of 1440 litres i.e. 1.2 l/s for 20 minutes.

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There are significant system cost and energy efficiency benefits of the Figure 8 system compared to the Figure 7 system. These are summarised as follows:

• The primary flow rate is halved resulting in smaller primary pipes and pumps

• The primary pump power is halved

• The required buffer vessel / thermal store volume is halved

• The secondary flow rate is reduced by a third resulting in smaller secondary pipes and pumps.

• The secondary pump power is reduced by a third

• Heat losses from secondary distribution pipes are reduced (due to smaller pipes operating at lower temperatures)

• There is the potential to add an auxiliary heat source such as a gas condensing boiler in the secondary return (i.e. feature E in Figure 1) where it can operate in full condensing mode by pre-heating the water from 40°C to 55°C.

These benefits will have a major impact on life cycle costing calculations for the alternative systems. In general, the energy and cost saving benefits increase if secondary flow temperatures are reduced and / or secondary ‘delta T’ values are increased. The only cost penalty involved is the fact that terminal units may need to be larger in size in order to compensate for operating at lower flow temperature and larger ‘delta T’. However, in modern buildings where localised heating loads are relatively small, the implications of having to install marginally larger terminal units is far outweighed in cost and energy terms by the benefits listed above.

3. SECONDARY CIRCUIT DESIGN

It is essential that secondary circuits dissipate heat effectively from the circulating water. This will enable the designer to maximise the design ‘delta T’ value and help to ensure that this is maintained (or exceeded) under all operating condition. If the return water should begin to increase in temperature under certain operating conditions, this may severely disrupt the operation of the system by:

• Preventing any auxiliary heat sources located in the secondary return from operating;

• Enabling high temperature water to enter the bottom of the buffer vessel / thermal store thereby disrupting the stratification layers within the vessel.

If the secondary distribution systems are to maintain consistently low return temperatures under all operating conditions, then careful consideration needs to be given to the elimination of any unwanted routes by which flow water can find its way directly back into return pipes. The following sub-sections explain the main secondary circuit design considerations for achieving and maintaining optimal ‘delta T’ conditions.

Variable flow systems

In order to gain any control over system ‘delta T’, the heating system must have a variable flow with 2 port control valves. This is in contrast to a constant flow system utilising constant speed pumps and 3 or 4 port control valves.

Constant flow systems operate at a fixed pump speed and flow rate throughout the year, regardless of system load. When zones are satisfied, heat output is reduced by diverting flow water through a by-pass which sends it straight

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back to the heat source. This inevitably causes return temperatures to increase rapidly towards the flow temperature thereby reducing the ‘delta T’.

By comparison, variable flow systems respond to part load conditions by throttling and reducing flows through terminal devices. At the same time, pump speed reduces simultaneously to suit the reduced flow requirement. The benefits of variable flow systems over constant flow systems are therefore:

• Reduced pump energy consumption – pump laws predict that each percentage reduction in pump speed produces a pump energy reduction of the same percentage cubed;

• Larger ‘delta T’ and lower return temperatures - when the flow rate through a terminal device is reduced, the water has more time to cool meaning that it returns to the heat source at a lower value.

Variable speed pumping is therefore critical to any ‘delta T’ design.

Pump speed control

Proper control of secondary pump speed will reduce the temperature increase across each pump and minimise pump energy consumption.

There are three common ways of controlling pump speed as described in the following bullet points.

• Constant pressure method - pump speed is controlled such that the pressure differential across the pump is maintained at a constant value equivalent to the pressure differential around the system at the maximum flow rate.

• Proportional method - pump speed is controlled such that the pressure differential across the pump reduces in proportion to flow rate towards a pre-selected value, typically equal to approximately 50% of the pressure differential around the system at the maximum flow rate.

• Remote sensor method - pump speed is controlled such that the pressure differential across the pump reduces towards the pressure differential across the most remote differential pressure controlled sub-branches, (or alternatively, at a point in the system where the differential pressure is approximately two thirds of the full load pump differential pressure).

These methods of control are illustrated by the pump and system curves illustrated in Figure 9 - 10 and Figure 11. For each method of control, the diagrams show the typical change between maximum load and the 50% part load condition.

It can be seen that the remote sensor option produces the biggest reduction between full load and part load operating speeds (i.e. between the red and red dotted lines). Since the percentage pump energy saving is equivalent to the percentage reduction in speed cubed, the remote sensor option provides by far the biggest saving in pump energy consumption.

This means that for larger systems, the aim should always be to control pump speed based on the signal from a remote differential pressure sensor located in the system. As a rule of thumb, in systems with a uniform load pattern, the sensor can be located at a point where the differential pressure is approximately two thirds of the maximum pump pressure differential. In systems where the load pattern is not uniform (i.e. the load is likely to shift between different system extremities) multiple sensors should be installed at each extremity to ensure that sufficient pressure and hence flow is available under all conditions.

Table 2 summarises the alternative methods of pump speed control and the applications for which they are best suited.

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Figure 9: Pump speed controlled to maintain constant pressure

Figure 10: Proportional control of pump speed

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Figure 11: Remote sensor control of pump speed.

Pump speed control method Suitable applications

Constant pressure control Primary circuits feeding multiple heat sources.

Proportional control Secondary circuits with total system pressure losses of less than 100kPa at maximum load.

Remote sensor control Secondary circuits with total system pressure losses of more than 100kPa at maximum load.

Table 2: Pump speed control options and applications

Terminal device selection / sizing

Terminal devices emit heat into occupied spaces. In general, the terminals best suited to ‘delta T’ designs are those that can work with low flow temperatures and generate large temperature differentials. The ability to dissipate heat is crucial to the performance of the system - an ideal terminal would be one that returns water at ambient air temperature having lost all of its heat to the surrounding space.

Of the various options, under-floor heating comes closest to meeting these requirements. With it’s large heat emission areas, flow temperatures can be maintained at 30-40°C so that low return temperatures can be easily achieved.

Other types of passive heat emitters such as radiant panels and radiators can also be made to conform to ‘delta T’ designs. For these types of emitters, design flow temperatures in the range 50-70°C can be selected enabling return temperatures in the range 20-40°C. Furthermore, circuits can be provided with weather compensation whereby the flow temperature is reduced as the external ambient temperature increases (this is discussed below under “Variable temperature circuits”).

Forced convection terminal units such as air handling units or fan coil units are more sensitive to flow variations, and the control of flow and return temperatures therefore requires more care. For these types of terminals, manufacturers may recommend that flow temperatures are maintained constant and design ‘delta T’ values do not exceed 20°C. Nevertheless, during selection, the aim should be to reduce the design flow temperature to its lowest possible value (typically in the range 55-70°C) in order to generate return temperatures in the range 35-50°C.

Outputs from forced convection terminal devices in variable flow systems are typically modulated by two port control valves. In such cases it is essential that these two port control valves are able to achieve accurate modulating control over heating water flow rates, otherwise return temperatures may not respond as intended. For example, if a 2 port control valve is ineffective at reducing flow until it is nearly closed, then the return water temperature will increase

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above its intended design value simply because the flow is not being slowed sufficiently by the control valve. This highlights the importance of selecting control valves with the correct “control characteristic” and with adequate “authority”. These terms are explained in the following sub-sections.

Underfloor Heating (UFH)

The ideal terminal unit is one that returns heating water at the ambient air temperature, having lost all of its heat to the surrounding space. Underfloor heating (UFH) comes closest to achieving this due to the large surface area of the heat emitter. Flow temperatures can be maintained at relatively low values resulting in return temperatures of around 30°C and less.

The ideal UFH heating system should achieve a high rate of heat dissipation enabling it to operate at a high flow rate with a smaller ‘delta T’ as opposed to low dissipation UFH which operates with a low flow rate and higher ‘delta T’. High rates of heat dissipation enable systems to be more reactive and operate at lower temperatures.

Figure 12: Thermal image comparison between Low and High Dissipation Underfloor Heating Systems

Figure 13: Performance data comparing Low and High Dissipation Underfloor Heating Systems

UFH Floor Temperature Distribution

Another challenge encountered with low dissipation UFH is uneven temperature distribution across the floor surface. A circuit with a low flow rate and high ‘delta T’ dissipates a greater proportion of the heat energy in the region of floor surrounding the start of the UFH circuit, leaving the area furthest from the manifold distinctly cooler to the touch. This temperature differential will create regions of hot and cold over the surface, which is not ideal from a comfort or control perspective. The design of the pipe layout can help to mitigate this problem with the counter-flow spiral being the most effective.

High dissipation UFH further reduces this problem by flowing at a higher rate and a lower ‘delta T’. The temperature differential between the start and end of the circuit is minimised resulting in a floor surface which radiates thermal energy evenly.

The normal objective of 70 / 40 is to increase the ‘delta T’ as much as possible to reduce the energy losses in the distribution pipework and achieve return temperatures which are as low as possible. However, in the case of UFH it is necessary to achieve the opposite and reduce the ‘delta T’ by increasing the flow rate. UFH pipework must be considered as a terminal unit, where “distribution losses” (or the output of the unit) need to be maximised. A reduction in ‘delta T’ in this case has the effect of increased reactivity and improved comfort for the user.

The importance of having high dissipation UFH is highlighted as this is only achievable if the heat taken from the pipes is dissipated rapidly, otherwise return temperatures will be elevated and the circuit becomes detrimental to ‘delta T’ design.

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UFH - Infra-red Floor Sensor

One method of improving underfl oor responsiveness is to monitor the temperature of the fl oor surface directly by infra-red fl oor surface temperature controllers, ensuring it remains inside the limits imposed by the BS EN1264-3 (29°C in occupied areas, 33°C in bath and shower rooms and 35°C in peripheral areas).

Figure 14: Infra-Red Floor Temperature Sensor

UFH - Heat Interface Units

In the multi-occupancy residential sector, including weather compensation within each HIU is usually cost prohibitive. Hence the UFH fl ow temperature is typically fi xed for design conditions.

For such applications, high dissipation UFH systems offer value in terms of better reactivity and lower fl ow and return temperatures.

Figure 15: Circuit Diagram of HIU with DHW and UFH

UFH - Difutec

SAV offers a high dissipation underfl oor heating solution, for more information please refer to the Difutec UFH Design Guide.

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Control valve characteristic

In order to achieve effective modulating control of heat transfer across a forced convection (i.e. fan driven) heating or cooling device, the control valve must be selected to suit both the emitter and the circuit in which it is located. The aim is to achieve an acceptable flow control characteristic.

A valve’s control characteristic is the relationship between the flow through the valve and its degree of opening. The characteristic is a feature of the design of either the valve itself, or the valve and actuator combination.

Percentage Valve Opening

Perc

enta

ge

Flo

w R

ate

0% 100% 50%

50%

100%

On/Off

Linea

r

Equa

l

Perc

enta

ge

Figure 16: Typical valve control characteristics

Standard valve control characteristics include on / off, linear and equal percentage. These are illustrated graphically in Figure 16 and can be described as follows:

• An on / off characteristic causes a large increase in flow over the initial 10-20% of its opening, and then a small increase in flow over the rest of its travel. For this reason, it is sometimes referred to as a “fast opening” valve.

• A linear characteristic causes an increase in flow that is directly proportional to the degree of valve opening.

• With an equal percentage characteristic, flow capacity increases exponentially with valve travel. Equal increments of valve travel produce equal percentage changes to flow.

On / off characteristics are suitable for applications where heat transfer is insensitive to flow variations e.g. passive heating or cooling devices. Linear characteristics are suitable for valves with a regulating function such as double regulating valves.

For active forced convection heating and cooling terminal devices, the best solution is an equal percentage characteristic. This is because for these types of terminal device, the heating or cooling output stabilises as water flow increases until a point is reached where the output becomes relatively unresponsive to further increases in flow.

22


Therefore, in order to achieve good modulating control over the heating or cooling output from forced convection terminal devices, the control valve needs to be most effective over the first part of its travel. This is best achieved by the equal percentage characteristic.

This characteristic can be produced by the design of the valve’s internals. Alternatively, it can also be generated by the valve actuator’s controlling software, by providing sufficient control points (i.e. discrete valve settings) to give an equal percentage characteristic. Where this type of “characterised actuator” is adopted, it must be used in conjunction with an appropriate and compatible control valve. Random mixing and matching of valves and actuators may result in ineffective equal percentage control.

Valve authority

If a valve used for controlling flow has a low resistance relative to the circuit in which it is located then, regardless of its intended flow control characteristic, it would need to close a long way before it starts to have an impact on flow rate. Such a valve would struggle to achieve control and would be described as having “poor authority”.

Therefore, in order to achieve effective modulating control, a valve needs to be sized such that its pressure loss, when fully open, is a significant proportion of the total pressure loss of the circuit containing the valve. This ensures that flow rate responds as soon as the valve begins to close.

The ratio of the control valve pressure loss when fully open to the pressure loss in the controlled circuit is referred to as “valve authority”. If the pressure loss across the fully open valve = p

1 and the pressure loss across all other

components in the controlled circuit = p2, valve authority (N) can be defined as,

N = p1 / (p

1 + p

2)

The maximum possible authority is 1. This occurs when p2 is zero and the only pressure loss in the controlled circuit

is across the control valve itself.

Figure 17 shows how an equal percentage flow characteristic can be distorted by installing the valve in a circuit within which it has poor authority.

Figure 17: Effect of valve authority on an equal percentage flow characteristic

This shows that a valve authority of 1 provides a perfect equal percentage characteristic, whereas at 0.5 the control is generally accepted as adequate. Two-port control valves have relatively low pressure loss and typically it may be difficult to achieve an authority of 0.5 or above. In such cases, authorities of 0.25 to 0.3 may have to be accepted on some valves depending on the criticality of the application. At an authority of less than 0.2 the flow characteristic becomes so distorted that it more closely resembles a linear characteristic.

23


It follows from the formula for N, that in order to give control valves sufficient authority, the pressure losses in the circuit in which they control flow must be limited. This will make it possible to select a control valve with sufficient authority for modulating control. The best way to limit pressure differentials across pipe circuits is by the use of differential pressure control valves, as described in the next sub-section.

Differential pressure control valves (DPCVs)

Differential pressure control valves (DPCVs) are used to control and limit pressure variations. A DPCV contains a diaphragm that separates the upper / lower chambers of the valve, which is connected to a spring-loaded piston. The valve opens when differential pressure rises and closes as the differential pressure falls, remaining steady during periods of equilibrium. In variable flow circuits, DPCVs are typically used to maintain a constant pressure differential across a sub-branch, protecting downstream control valves from excessive pressures and neutralising the effects of pressure variation caused by other control valves in the circuit.

A typical situation is illustrated in Figure 18 which shows a heating circuit with a radiator thermostatic radiator valves (TRV). Without DPCV protection, then as an individual TRV begins to close, the pressure differentials across the remaining fully open radiator circuits could be expected to increase.

Pressureheldconstantby DPCV

Pressuregradient atfull load

Pressuregradient atpart load

Variation inpressureacross endterminalMin

Max

Figure 18: Potential for excess pressures and hence flows through terminal devices

This situation would lead to excess flow through the radiator units still with fully open TRVs and a consequent reduction in the system ‘delta T’ values. The action of room temperature controls cannot be relied on to correct this problem. Excess flows through radiators do not cause significant increases in heat output, and, it may therefore be quite some time before the excess flow causes an increase in room temperature sufficient to get the TRVs to throttle.

In order to maintain the system design ‘delta T’ value at part load, it is important to avoid large increases in differential pressure across terminal branches. This can be achieved simply and effectively by placing a DPCV across the flow / return connections local to the first radiator TRV, as shown in figure 18.

Typically, DPCVs on branches serving groups of radiators with TRVs should be sized to maintain a differential pressure of not more than 10kPa. Similarly, DPCVs on branches serving groups of terminal units with two port valves should be sized to limit pressure differential to no more than 1.5 times the design pressure loss across the highest resistance terminal unit branch (inclusive of its two port valve). This will limit the maximum flows through the terminals under part load conditions, thereby helping to maintain control of the system ‘delta T’.

24


DPCV Proportional Band

As DPCVs are not electronically compensated, deviations from their set point are inevitable. The proportional effect depends on the correlation between the controller valve’s degree of opening and the deviation between the controlled and set differential pressure. Furthermore, the deviation depends on the actual differential pressure across the control valve and the actual control setting.

Reference figure 19 it can be seen that the Danfoss AVPL DPCV has a narrow proportional band of ± 1…3 kPa, depending on the actual differential pressure and setting. In this example a 3 kPa “P” band means the valves controlled differential pressure can fluctuate between 15 and 18 kPa across its flow and differential pressure range.

NB: Correct DPCV KVS selection is vital for achieving a narrow “P” band.

Settings

Figure 19: Danfoss AVPL DPCV “P” Band

Alternative solutions to the problem of excess pressures across terminal branches are described in the BSRIA Guide BG12/2011 Energy Efficient Pumping. This guide recommends the following options:

• The use of DPCVs as part of valve modules (see SAV FloCon range). The action of DPCVs to maintain a narrow differential pressure range constant across circuits fed from a manifold arrangement will help to ensure that flows through individual terminal units remain within acceptable limits regardless of closures in adjacent circuits.

Pump minimum flow rate

When in service, a pump requires a minimum flow rate. Without such flow, most of the power generated by the pump motor is converted to heat, with undesirable consequences.

In variable flow systems, by-passes should be provided to ensure that the minimum pump flow requirement is always satisfied. To help maintain design ‘delta T’ values, the flow in the system at minimum load must be kept as low as possible. An assessment should therefore be made of the minimum acceptable flow through the pumps and this value should be used as the target to be achieved under minimum load conditions.

The pump speed turndown ratio, and hence the minimum flow rate that can be tolerated, is dependent on the type of pump. As a general rule, the pump speed turndown ratio should not drop below 10%.

With glandless motor pumps, some water flows through the pump casing in order to cool the motor bearings. However, if the pump flow rate is too low, this would lead to overheating of the bearings, producing an alarm condition and shut down. As a rule of thumb, glandless pumps should not be operated at less than 7.5% of their nominal flow rate where the nominal flow rate can be estimated as approximately 75% of the manufacturer’s maximum flow rate value. See Figure 20.

25


Figure 20: Minimum flow rate for canned rotor pumps

For pumps where the motor is separated by a drive shaft, the minimum permissible flow rate is dictated by what is considered to be an acceptable temperature increase across the pump. At near zero flow, the majority of power generated by the motor is converted to heat, causing the water in the pump to increase in temperature. The maximum possible increase in temperature can be calculated from the equation:

DT = Pzero

/ (cp q

m)

Where: DT = temperature increase of the liquid as it passes through the pump (K)

Pzero

= pump power at zero flow (W)

cp = specific heat capacity of fluid (which for water is 4187 J/kgK)

qm = mass flow rate of water (kg/s).

The pump power available to heat the liquid at zero or near zero flow can be determined from the pump manufacturer’s published data. A typical manufacturer’s pump power curves are shown in Figure 21. The value of P

zero is the pump power at zero flow; it is assumed that when the pump is operating at its minimum load condition

i.e. at minimum speed, its power consumption will be approximately equal to this value.

Hence, using the above equation, a value for the maximum temperature increase across the pump can be calculated. The designer must then decide what level of temperature increase is acceptable.

For heating systems, the temperature increase must be limited to avoid increasing water temperature to a point where a high temperature alarm condition might be triggered. A temperature increase of 2-3°C is usually acceptable in heating systems, enabling minimum flow rates in the range 5 - 10%.

It can be seen from Figure 21 that the lower the pump pressure at minimum load, the lower the value of Pzero

and hence the smaller the temperature increase across the pump. It is therefore essential that secondary pump speed is controlled in such a way that pump pressure can drop to the lowest possible value. This is best achieved by remote sensor control of pump speed.

26


Figure 21: Pump power consumption at different operating speeds

System by-passes

System by-passes can be thought of as any feature in the system which allows flow water to return to the heat source, without having been cooled by passing through some form of heat exchanger. Since by-passed water is uncooled water, it will have the effect of increasing return water temperatures. If the optimal design ‘delta T’ value is to be maintained, it is essential that secondary system by-passes should be minimised.

However, some by-passes are essential for satisfactory system operation. By-passes are commonly included in secondary distribution systems in order to:

• Ensure that system flow does not drop below the pump minimum flow value (as explained in the preceding sub-section);

• Make allowance for a future heating load to be installed at a later date;

• Maintain pipes in a “live” condition ready for heating demand;

• Maintain the circulation of water treatment chemicals to system extremities.

To best satisfy the last two bullet points on this list, by-passes should always be installed at system extremities i.e. at the ends of branches. The three best by-pass options are as follows:

• Variable flow. A variable flow by-pass is shown in Figure 22. The by-pass includes a fixed orifice double regulating valve for flow regulation and flow measurement. Because there is no control of the differential pressure across the by-pass, the flow through the by-pass will vary under different load conditions. The by-passes should be sized such that their summated flows are equal to the target minimum flow rate value.

• Constant flow. A constant flow by-pass is shown in Figure 23. This type of by-pass is particularly suited to systems feeding terminal devices fitted with PICVs. The by-passes should be sized such that their summated flows are equal to the target minimum flow rate value. Constant flow regulators should be included in each by-pass to ensure that under all operation conditions the target flow rate is maintained.

• Integral. An integral by-pass is shown in Figure 24. Integral by-passes are those included within the control valves themselves. Examples are 3 or 4 port valves that divert flow through a by-pass when the load is satisfied.

27


To keep pipes ‘live’ as far as possible, multiple low flow by-passes should be installed at system extremities. The total amount of by-passing flow should be in the range of 5-10% of the total design flow (having checked that this is not less than the required pump minimum flow rate). If this leaves some parts of the system without regular flow, the circulation of water treatment chemicals can still be achieved by exercising the valves periodically (e.g. cycling all two port valves fully open for say, 10 minutes in every 48 hour period during the non-heating season).

Variable flowby-pass

Figure 22: Typical circuit with variable flow by-pass at end of run

Constant flowby-pass

Figure 23: typical circuit with constant flow by-pass at end of run

28


Integralby-pass(4PV)

Figure 24: Typical circuit with integral (4PV) by-pass at end of run

Variable temperature circuits

In order to achieve good ‘delta T’ design, terminal device flow and return temperatures must be as low as possible. The use of variable temperature circuits (as opposed to constant temperature circuits) is critical to achieving this objective.

Figure 25 shows constant and variable temperature secondary circuits. The key difference is that in the case of variable temperature circuits, a three port valve is used to mix some of the returning water back into the flow water thereby reducing its temperature. The degree of mixing will determine the flow temperature. Constant temperature circuits have no such provision to control flow temperature.

CONSTANTTEMPERATUREVARIABLE

TEMPERATURE

PRIMARYHEADER

Figure 25: Constant and variable temperature secondary circuit layouts

The choice of either constant or variable temperature circuits has traditionally taken into account of the type of terminal devices to be fed from the system and how best to control their heat output. Natural convectors such as radiators or under-floor heating tend to be fed from variable temperature circuits because they are relatively insensitive to flow variations. Heat output from these emitters is therefore best controlled by varying the flow temperature. On the other hand, forced convection heat emitters such as fan coil units have traditionally been fed from constant temperature circuits. This is because their heat output is sensitive to flow, which can be varied in turn by modulating control valves.

However, regardless of the type of terminal device, in ‘delta T’ designs variable temperature circuits are essential if optimal flow and return temperature values are to be maintained.

This is because a variable temperature circuit enables the secondary flow and return temperature to be reduced to their lowest possible values, thereby maximising system efficiency. For example, the benefits accruing from the system shown in Figure 8 of this guide, relative to that shown in Figure 7 are only possible because the flow temperatures in the secondary circuits can be reduced to values that are lower than those required in the primary flow pipes.

29


‘Delta T’ controllers - Weather compensation and return temperature regulator

It is important that every opportunity be taken to reduce the heating system flow temperature whenever possible. Intelligent ‘delta T’ controllers can achieve this through weather compensation alone or a combination of weather compensation and return temperature monitoring. This will have the following benefits:

• Heat emissions from distribution pipes will be reduced thereby avoiding energy wastage. This is particularly relevant for large district heating systems where pipes are run externally and are exposed to low surrounding air temperatures.

• The potential for high return temperatures under low load operating conditions will be mitigated. Return temperatures tend to increase towards the flow temperature under low load conditions because the proportion of flow circulating through by-passes increases relative to the overall flow through the system.

• Heat outputs from terminals will become easier to control, resulting in less risk of over-heating occupied spaces.

To some extent the heating system flow temperature will be determined by the application. For example, a system serving under-floor heating can be run at a relatively low flow temperature (35°-40°C). However, heating systems serving hot water heaters may require flow temperatures in the range of 50°C, in order to generate hot water at 40°C. Where it makes sense, pipe systems should be split so that terminal devices with different temperature requirements are fed from separate secondary circuits.

‘Delta T’ controllers can be provided as stand-alone units by SAV (UK) Ltd.

Self standing ‘delta T’ controllers

In situations where different types of terminal device are to be fed from the same heating system, it may be more cost effective and user friendly to provide localised self-standing ‘delta T’ controllers to manage variable temperature circuits. A typical configuration is shown in Figure 26.

Delta ‘T’ Controller

Variabletemperaturecircuit

Heatingmains


MODULATINGCV

DPCV

RECIRCULATIONLINE

Figure 26: Localised weather compensated variable temperature UFH sub-circuit feeding off a higher temperature radiator circuit.

It can be seen that this solution requires a separate pump, sized to cope with the flow through the variable temperature circuit.

If the flow temperature sensor goes above set point (e.g. 35ºC), the ‘delta T’ controller instructs the 2-port valve to close, thus promoting movement through the recirculation line. Similarly, if the return temperature exceeds set point (e.g. 30ºC), the 2-port valve also closes. Only 1 temperature excess signal is required to actuate the 2-port valve, thus providing control of flow and return temperatures.

30


‘Delta T’ Monitoring - Continuous digital monitoring of HVAC systems

SAV’s Flocon Watchman commissioning modules enable continuous digital monitoring of flow rate and return temperature for individual zones and terminals within building HVAC systems.

The combination of a flow meter on the main supply pipe with two port control valves on each of the terminal branches means that precise flow rate information can readily be obtained. By closing the two port control valve on any particular terminal branch, even ultra low flow rates can be derived from the drop in total flow through the meter. This is known as the subtraction method.

Digitally monitoring the return water temperature from individual terminals enables a full soft landings approach whereby the theoretical return temperature predictions are tested under operating conditions. If return temperatures are not as predicted producing a shortfall in ‘delta T’, this will be highlighted by the BMS. This provides the information to enable fine tuning of the system.

Each module also includes an energy meter that measures and records the energy consumptions of up to 7 terminal units at a time. This means that if a building is sub-divided into multiple tenancies, each lettable area can be metered individually and the tenants billed accordingly.

‘DELTA T’ ENERGY METER CALCULATION METHODOLOGY

The following example shows the methodology for calculating the average ‘delta T’ between two points in time using energy meter data.

Reading 1

m3 MWh

Reading 2

Reading 1

Reading 2

m3 00120 MWh

Cubic Meter Consumption in period: 1340 – 780 = 560 m3

Energy Consumption in period: 120 – 101 = 19 MWh

‘Delta T’ calculated as follows:

MWh x 860 = heating in °C m3

The 860 is a constant, and is defined as the quantity of m3

of water that will be heated by 1ºC by 1 MWh

So in this example the ‘delta T’ calculation is as follows:

19 x 860 = 29°C 560

The average ‘delta T’ in this example is therefore 29ºC

870

01340

00101

ChilledWater

B

A

A = FloCon BMS Interface

B = MID Certified Digital Flow and Energy Meter - Chilled Water

C = MID Certified Digital Flow and Energy Meter - Heating Water

C1, C2, C3, C4 = Return Temperature - Chilled Terminal Unit

H1, H2, H3, H4 = Return Temperature - Heating Terminal Unit

C

HeatingWater

FloCon BMSInterface

C1 C2 C3 C4 H1 H2 H3 H4

Figure 27: Schematic of FloCon module for heating and cooling duties

31


Hot water heating and Heat Interface Units

Secondary circuits feeding direct heated hot water storage cylinders are not always compatible in systems that require low return temperatures and high ‘delta T’ values. Hot water cylinders are usually designed to heat water to 60°C, the minimum temperature required to kill legionella bacteria. To heat the entire contents of a cylinder to this temperature within a reasonable time span requires a high flow temperature and low ‘delta T’, resulting in return temperatures equal to or greater than 60°C.

As an alternative to storage cylinders, hot water can be generated by the use of plate heat exchangers that can provide much higher heat dissipation than traditional cylinders. This solution is available from SAV Systems in the form of packaged heat interface units (such as Danfoss FlatStations).

This type of unit provides instantaneous heating of the incoming domestic cold water by means of a plate heat exchanger. Heating water circulates on one side of the plate heat exchanger transferring heat to mains cold water passing through on the other side. The heated water can then be fed straight to the hot water taps, or to a hot water cylinder.

A typical heat interface unit (or FlatStation HIU) is shown in Figure 28. This identifies three valves (A, B & C) which are critical to the function of these units.

A. A DPCV must be installed in the circuit feeding the radiators. This enables the pressure differential across the radiator circuit to be set to be maintained at its optimum value - typically less than 10kPa for apartments and standard sized houses. This enables the radiator TRVs to control more effectively and will help to maintain a consistent balance of flows through the radiators. During commissioning, the DPCV should be set to maintain the required design flow rate through the radiator circuit with radiators sized at the optimal ‘delta T’. Because the DPCV is a self-acting valve it will modulate automatically to maintain a constant pressure differential across the circuit regardless of changes in pump speed or valve closures elsewhere in the system.

B. A fast acting temperature control valve is required which simultaneously regulates both the heating water flow and the flow of cold water into the plate heat exchanger. This is to maintain constant domestic hot water temperature, regardless of variations in heating or cold water conditions. This type of valve maximises the ‘delta T’ across the unit during hot water production.

Note: This type of valve is critical to ‘delta T’ design. Heat interface units that rely on slow acting control valves will inevitably produce variations in domestic hot water temperatures and result in unwanted high return temperatures.

C. A DPCV must be located in the heating circuit feeding the plate heat exchanger, to maintain to a constant pressure across the heating side of the exchanger at times of domestic hot water demand. Because the DPCV is self-acting, it modulates automatically to maintain a constant pressure differential, regardless of changes in pump speed or valve closures elsewhere in the system.

Figure 28: Danfoss FlatStation Heat Interface Unit for instantaneous indirect

hot water heating and direct space heating

32


Since hot water is heated instantaneously in a FlatStation HIU, there is no necessity for hot water storage. Due to the high rate of heat transfer into incoming cold water, heating water entering at 60-70°C can be consistently cooled to 25°C during draw-off, making these units ideal for ‘delta T’ designs.

It is important that low primary return temperatures be maintained at part load, regardless of fluctuations in demand or primary pressures.

The relationship between hot water demand and system operating temperatures using FlatStation HIUs is illustrated in Figure 29. As can be seen from the performance chart, the primary return temperature is consistently kept below 25°C as recommended by CIBSE AM12 2013 and as required by Greater London Authority’s District Heating Manual for London 2013.

When there is no draw-off of hot water, the self-acting control valves inside FlatStation HIUs throttle the heating flow to a minimum, thereby causing the pump to reduce its speed and power consumption. Instead of being closed off completely, the flow through the unit is finely controlled to an amount just sufficient to maintain a heating return temperature 8°C below the set hot water tapping temperature. This feature prevents the pipes feeding the unit from becoming dead legs and ensures that hot water is available as soon as taps are opened. It also keeps return temperatures as low as possible.

1m 3m 4m 5m 6m2m

Figure 29: Domestic hot water demand and operating temperatures using Danfoss FlatStation HIU

Where storage of hot water is a requirement (as might be appropriate in a centralised hot water system for a school or similar building) a plate heat exchanger can be used to control system ‘delta T’ whilst providing heat to a storage cylinder. Figure 30 shows a typical configuration.

In order to generate domestic hot water at 60°C (the recommended temperature for eliminating legionella), the heating water flow must be at a temperature higher than this. Hot water is pumped to a storage cylinder which fills from the top down. A specialised fast acting pressure independent control valve (DHW-PICV) is used to control the flow of heating water through the plate heat exchanger. This valve closes when the hot water recirculation increases above its set point value (indicating that the cylinder is full). As hot water is drawn from the system, it is replaced by cold water resulting in the opening again of the DHW-PICV to start generating more hot water.

Hence, heat input is only provided when hot water generation is required, enabling the heating system temperature to remain low at 40°C or less.

33


Figure 30: Plate heat exchanger for heating of hot water in cylinders

For systems with multiple distributed outlets, hot water circulation can be provided with either instantaneous or storage type hot water solutions, as indicated in Figures 26 and 29. The design of systems incorporating these units is explained in the SAV FlatStation Design Guide 2014.

34


4. ‘DELTA T’ COMMISSIONING AND SOFT LANDING

Commissioning

Proper commissioning of the system is essential if the ‘delta T’ design is to be effective. It is not sufficient to simply check that all design flow rates are achievable. There must also be a check of system operating temperatures.

Proper commissioning of ‘delta T’ designs should comply with the requirements of CIBSE Commissioning Code W: Water Distribution Systems 2010 and BSRIA Guide BG2 / 2010 Commissioning Water Systems. However, particular attention should also be paid to the following issues:

• Design flow rates should be tested under both maximum and minimum load conditions. It is particularly important to prove that the specified minimum flow rate is achieved when the system is at minimum load (i.e. with all control valves fully closed). This should normally be a value in the region of 5-10% of the maximum.

• Pump pressure differentials should be measured at both maximum and minimum load conditions. This will demonstrate whether the specified method of pump speed control has been properly implemented.

• Variable flow by-passes should be commissioned at minimum load (all control valves fully closed). If the specified by-pass flow rate were set with the system operating at full load, the flows through these by-passes would inevitably increase in an uncontrolled way as control valves close and the differential pressure across the by-pass increases.

• The differential pressures controlled by DPCVs should be measured and recorded to ensure that the values are in line with design and that there is no risk of excessive flows through terminal branches.

Having ensured that commissioned design flow rates at maximum and minimum load conditions are correct, the system return temperature should be monitored to ensure that its specified design value is being maintained. If the ‘delta T’ design is successful, the returning water temperature should remain at or below the target return temperature under all operating conditions.

Soft Landings

The refinement of building systems operation to optimise system temperatures should continue after construction hand-over. The proper functioning of the system is strongly reliant on maintaining ‘delta T’ values. There is always the potential for the system ‘delta T’ to be squeezed by a poorly functioning control valve or an open by-pass. For this reason, systems should be monitored during their operation and any reductions in ‘delta T’ investigated and corrected.

Please contact SAV Systems for more detailed information on energy metering systems or any other aspects of this design guide.

35


districtenergy.danfoss.com

Results and experiences from a 2 - year study with measurements on a low - temperature DH system for low energy buildingsChristian Holm Christiansen1, Alessandro Dalla Rosa2, Marek Brand2, Peter Kaarup Olsen3, Jan Eric Thorsen4 DHC13, the 13th International Symposium on District Heating and CoolingSeptember 3rd to September 4th, 2012, Copenhagen, Denmark

1 Danish Technological Institute, Gregersensvej 1, DK-2630 Taastrup, Denmark 2 Technical University of Denmark, Dept. of Civil Engineering, Brovej, DK- 2800 Kgs. Lyngby, Denmark 3 COWI A/S, Parallelvej 2, 2800 DK-Kgs. Lyngby, Denmark4 Danfoss District Energy, Nordborg, Denmark, +45 7488 4494, [email protected]

Technical paper

MAKING MODERN LIVING POSSIBLE

districtenergy.danfoss.com

Results and experiences from a 2 - year study with measurements on a low - temperature DH system for low energy buildingsChristian Holm Christiansen1, Alessandro Dalla Rosa2, Marek Brand2, Peter Kaarup Olsen3, Jan Eric Thorsen4 DHC13, the 13th International Symposium on District Heating and CoolingSeptember 3rd to September 4th, 2012, Copenhagen, Denmark

1 Danish Technological Institute, Gregersensvej 1, DK-2630 Taastrup, Denmark 2 Technical University of Denmark, Dept. of Civil Engineering, Brovej, DK- 2800 Kgs. Lyngby, Denmark 3 COWI A/S, Parallelvej 2, 2800 DK-Kgs. Lyngby, Denmark4 Danfoss District Energy, Nordborg, Denmark, +45 7488 4494, [email protected]

Technical paper

MAKING MODERN LIVING POSSIBLE

2 Danfoss District Energy

Introduction

An innovative low-temperature District Heating ( DH ) system for low - energy buildings that operates with supply tem-peratures slightly above 50 °C has been successfully put in operation in 2010. This paper presents the results of a 2 - year study with detailed measurements of a low heat density area with 40 low - energy terraced houses and a communal build-ing in Lystrup, Denmark, see table 1 and figure 1. The project dealt with the inte-gration of sustainable solutions both for the end-user side and the energy supply side and aimed to:• Demonstrate the operation and energy

demand of DH applied to low - energy buildings and that the heat loss in the network can be maintained below 15 – 20 % of the total delivered heat.

TECHNICAL PAPERResults and experiences from a 2 - year study with measurements on a low - temperature DH system for low - energy buildings

Author(s)

Jan Eric Thorsen,Danfoss District Energy, Nordborg, Denmark, +45 7488 4494 · [email protected]

Christian Holm Christiansen,Danish Technological Institute, Gregersensvej 1, DK-2630 Taastrup, Denmark

Alessandro Dalla Rosa,Marek Brand,Technical University of Denmark, Dept. of Civil Engineering, Brovej, DK- 2800 Kgs. Lyngby, Denmark

Peter Kaarup Olsen,COWI A/S, Parallelvej 2, 2800 DK-Kgs. Lyngby, Denmark

A new low-temperature district heating system for low-energy buildings that operates with supply temperatures slightly above 50 °C was presented at the 11th International Symposium of District Heating and Cooling in 2008; the design includes newly developed substations and efficient distribution pipes, resulting in reduction of heat losses up to 75 % compared to traditional layouts. Since then, the first area using the new system has successfully been put in operation. This paper presents the results of a 2 - year study with detailed measurements of a low heat density area with 40 low - energy terraced houses in Denmark. The investigations include the determination of the heat losses from the distribution network, the pumping electricity consumption, the user behavior in terms of indoor temperature and domestic hot water consumption as well as detailed simultaneity factors to be used for network design. Moreover, the paper presents solutions for using the return water of existing networks to supply district heating to newly built areas and summarizes in general on how to integrate low - energy houses and district heating systems. Finally, it points to the potential of integrating low - temperature district heating systems in existing buildings as an effective solution towards energy - sustainability in the heating sector.

Project information

owner Housing association BF Ringgården

year of construction 2008 – 2010

site area [ ha ] 1.7

building units ( residential ) 40 terraced houses

residents Seniors, young families

number of residents 92 ( estimated )

building units ( teritary ) 1 communal building

heated area [ m2 ] 4115

plot ratio 1 0.241 built floor area / site area

TABLE 1: Basic information on the project

Project information

owner Housing association BF Ringgården

year of construction 2008 – 2010

site area [ ha ] 1.7

building units ( residential ) 40 terraced houses

residents Seniors, young families

number of residents 92 ( estimated )

building units ( teritary ) 1 communal building

heated area [ m2 ] 4115

plot ratio 1 0.241 built floor area / site area

TABLE 1: Basic information on the project

Extracts - Danfoss Lystrup Technical Paper

50 30

36


Heat demand

The building installations, in terms of heating system, consist of a combination of radiators – based on design supply/return / room temperature of 55 / 25 / 20 °C

DHW is prepared by one of the low - tem-

Figure 2: Sketch of the DHSU and IHEU principles of DHW production

DHSU principleDH supply

DH return

120 lDHW

DCW

IHEU principleDH supply

DH return

DHW

DCW

thermostatic bypass

perature DHW systems described in [ 1 ], [ 2 ]: the low - temperature Instantaneous Heat Exchanger Unit ( IHEU ) and the low-temperature District Heating Storage Unit ( DHSU 2.

The layouts of the DHW distribution pipes

carefully designed, so that there is a sep-

and the length of the pipe is minimized, 3. Consequently, the water con-

tent in each DHW supply line, including the volume in the secondary side of the DHW heat exchanger, is kept to a mini-mum and it is below 3 liter: this is the max-imum allowable water content that assures safety in relation to the Legionella risk, even without any treatments ( ther-mal, UV - rays or chemical ), according to the German guidelines for DHW systems ( DVGW, W551 ).

Figure 4: Sketch of the low - temperature network with the location of the meters. DH is delivered from the utility Lystrup Fjernvarmeto the pumping station

Instantaneous Exchanger Unit (IHEU) flow rate meter: 0.01 – 0.38 m 3 / h

Pumping station flow rate meter: 0.8 – 12 m 3 / h

flow rate meter (street): 0.26 – 3.4 m 3 / h

District Heating Storage Unit (DHSU) flow rate meter: 0.01 – 0.09 m 3 / h



perature DHW systems described in [ 1 ], [ 2 ]: the low - temperature Instantaneous Heat Exchanger Unit ( IHEU ) and the low-temperature District Heating Storage Unit ( DHSU ), see figure 2.

The layouts of the DHW distribution pipes and the floor plan of the dwellings were carefully designed, so that there is a sep-arate pipe supplying each DHW fixture and the length of the pipe is minimized, see figure 3. Consequently, the water con-tent in each DHW supply line, including the volume in the secondary side of the DHW heat exchanger, is kept to a mini-mum and it is below 3 liter: this is the max-imum allowable water content that assures safety in relation to the Legionella risk, even without any treatments ( ther-mal, UV - rays or chemical ), according to the German guidelines for DHW systems ( DVGW, W551 ).

Heat distribution network

A sketch of the DH network with the loca-tion of flow meters for monitoring is seen in figure 4. Besides the normal end - user heat meters and the main meter placed at the pumping station two additional meters are placed at the end of two dif-ferent streets. One meter is measuring a part of the network with 11 DHSU's, the other is measuring on a network part with 11 IHEU's.

Network dimensioning

The network consists of flexible plastic twin pipes for dimensions up to DN32 and of steel twin pipes for larger dimensions. Heat loss coefficients are calculated according to [ 2 ] and pipe manufacturer data, figure 4.The other assumptions for the design were:• Maximum pressure level: 10 bar. It is

reasonable to design the network according to the maximum hydraulic load that can be withstood by the distribution pipeline; in this case the limit is drawn by the plastic service pipes, which requires pressure levels below 10 bar. In fact the pipeline systems must by regulations withstand pressures 1.2 – 1.5 times the nominal value.

• Thermostatic by-pass valves of IHEU's set to 40 °C, in the customer's substation at the end of each street line

FIGURE 3: Sketch of the floor plans with the layout of the DHW distribution pipelines. Type C1 ( left ), type C2 ( right )

FIGURE 4: Sketch of the low - temperature network with the location of the meters. DH is delivered from the utility Lystrup Fjernvarmeto the pumping station

block number total size [ m2 ]number of dwellings

type c1 type c2

1 771 5 3

2 727 2 5

3 594 3 3

4 528 1 4

5* 479 1 2

6 484 3 2

7 532 6 0

* including the communal building, A = 170 m2

TABLE 3: Type and floor area of the buildings

substation DHW pipe

Instantaneous Exchanger Unit (IHEU) flow rate meter: 0.01 – 0.38 m3 / h

Pumping station flow rate meter: 0.8 – 12 m3 / h

flow rate meter (street): 0.26 – 3.4 m3 / h

District Heating Storage Unit (DHSU) flow rate meter: 0.01 – 0.09 m3 / h

Technical Paper Results and experiences…

and set to 35 °C, in all the other customers' substations.

• Design supply temperature from the mixing shunt: 55 °C; design return temperature: 25 °C.

• Maximum water velocity: 2.0 m / s; also in branch pipes.

37



5Danfoss District Energy

building design values. Further, the indoor temperature, as measured in the living room, was 2 – 4 °C above the design indoor temperature of 20 °C during the heating season, see figure 7. In [ 1 ] it was shown that 1 °C higher than expected room temperature can lead to 20 % higher SH demand in low - energy houses, so the room temperature alone, almost explains the high consumption.DHW consumption was measured to be 65 liter / ( day × house) in average. It is a low value, which is partly related to the number of occupants and their compo-sition. Based on an estimate of the num-ber of residents in the dwellings, it is assessed that the DHW use was equiva-lent to approx. 28 liter / ( day × person ). It should be noticed that the average cold water temperature was approx. 15 °C and the average DHW temperature was 40 – 45 °C, giving an average temperature difference of 25 – 30 °C in the first moni-toring period. According to [ 8 ], DHW use of 30 – 40 liter /(day × person) and a tem-perature difference of approx. 40 °C are typical values for Denmark. In the case study this would give an expected heat demand for DHW of 12 – 16 kWh / ( m2 × yr ). In average the DHW consumption mea-sured was 8 kWh/( m2 × yr ), which is less than the design value ( table 2 ). When added to the SH demand a total annual DH consumption of 5.8 MWh per house was found. Monitoring also demon-strated that DHW can be produced at temperatures of just 3 °C below the pri-mary supply temperature, e.g. 47 °C at a DH supply temperature of 50 °C as expected with the used substations.

Monitoring

An extensive monitoring program and data acquisition system was established; the measurements presented here are mainly from the first monitoring period conducted during the weeks 26 – 47, 2010 with the main meters and individual meters in 22 dwellings in place; in these dwellings both DH meter, meter for DHW and a room temperature sensor were installed. Late 2010, DH meters in the remaining 19 dwellings were connected and the monitoring has continued ever since and will go on until end of 2013.

Space heating and domestic hot water

Based on the measurements in the first monitoring period, a heat load vs out-door temperature curve was estab-lished, see figure 6. For IHEU the standby heat loss is about 25 W; for DHSU 80 W. It is seen that the average heat load dur-ing summer ( week 26 – 38 ) is higher than the standby losses which means that there has also been a SH demand in this period for some houses e.g. for floor heating of the bathroom.Based on the curve, the SH demand per dwelling was estimated to 5.1 MWh for the Danish reference year corresponding to about 51 kWh / ( m2 × yr ), 70 % higher SH demand than building regulation design value ( see table 3 ). It was not the purpose of the project group to look at the building „as - built” vs. designed. Analysis of the measurements also rather indicates that the reason should be found in user behavior.Floor heating during summer is a param-eter that is not taken into account in the

and set to 35 °C, in all the other customers' substations.

• Design supply temperature from the mixing shunt: 55 °C; design return temperature: 25 °C.

• Maximum water velocity: 2.0 m / s; also in branch pipes.

• The simultaneity factor was assumed to be 1.0 in case of DHSU, due to the low semi - constant flow the unit was designed for. The simultaneity factor for the IHEU was the traditional consumer dependent approach used in Denmark. Design loads are for DHSU: 3 kW, for IHEU: 32,3 kW.

• Minimum supply / return pressure difference at the end - user's substation: 0.3 bar.

Heat sources

The distribution network in this case study is a typical example of how a low-temperature DH scheme can be inte-grated in an existing network that has higher operating temperature. There are no heat sources on the site. The heat is provided directly from the medium - tem-perature DH utility Lystrup Fjernvarme. A pumping station and a mixing shunt are placed in the communal house. The pump is operated based on a pressure difference sensor placed at the critical point in the network. The mixing shunt is controlled by a return valve and a tem-perature sensor in the main supply pipe to the low-temperature network. The sys-tem is seen in Figure 5 together with the pressure line drawn from the pumping station to the end - user.

inner diameter

[ mm ]

[ W / ( m × K ) ]roughness

[ mm ]lenght

[ m ]

estimated cost in 2010 [ €/m ]

U11

=U22

U12

=U21

purchase total 1

Alx 14/14 - 110 10 0.05 0.035 0.02 123 47 162

Alx 20/20 - 110 15 0.065 0.037 0.02 221 56 166

Alx 26/26 - 125 20 0.071 0.049 0.02 155 67 207

Alx 32/32 - 125 26 0.088 0.053 0.02 130 78 211

Tws - DN 32 37.2 0.085 0.056 0.1 90 82 240

Tws - DN 40 43.1 0.099 0.053 0.1 32 88 246

Tws - DN 50 54.5 0.096 0.06 0.1 16 122 268

TABLE 4: Pipe specifications. Alx: Aluflex twin pipes; Tws: Steel twin pipes, series 2, diffusion barrier at the outer casing.



Operating temeratures

Lystrup Fjernvarme that supplies heat to the new low - temperature area is a medium - temperature DH system. DH is supplied with up to 80 °C during winter and down to 60 °C during summer. In fig-ure 8, the average weekly supply and return temperatures and heat load are seen for the 2 year monitoring period together with the shunted supply tem-perature. The maximum monitored heat load is 161.3 kW compared to maximum weekly average of 87.4 kW. The figure shows how the shunt has been adjusted during the period in order to get the low mixed supply temperature of just slightly above 50 °C. Further the result of trouble-shooting in individual building installa-tions has secured a low return tempera-ture. This is expected to be even lower after the local „boiler man” beginning 2012 has been provided a tool that updates him every week with return tem-perature and other relevant data of the individual houses. Based on the tool, he can guide the users towards better oper-ation of the substations which has already been beneficial.In the first monitoring period of 2010 the two different substation types were com-pared specifically. In the 11 homes with DHSU, the average return temperature was 39.4 °C in the weeks 26 – 47; in sum-mer – weeks 26 – 38 – the average return temperature was 43.6 °C. The high return temperature was primarily due to the malfunction of a single unit. The best per-forming DHSU registered a return tem-perature of 29 °C in summer. The 11 homes with IHEU, the average return tem-perature was 34.7 °C in the weeks 26 – 47; in summer – weeks 26 – 38 – the average return temperature was 40.3 °C. The high return temperature was primarily due to 2 substations, where the control valves were defected and allowed a relative large amount of water to flow uncooled to the return pipe. The best performing IHEU registered a return temperature of 26 °C in summer. Observing the same weeks, a year later, showed improved results as seen in figure 9 for the 11 IHEU's, even though the supply temperature had been reduced in the meantime.In general, the return temperature in the heating season ( week 39 – 47 ) was lower than during the summer, which confirms that the radiators delivers low return tem-peratures ( 28 – 33 °C ). This occurred

FIGURE 5: Simplified pressure line / temperature diagram of the mixing shunt during typical operating conditions

FIGURE 6: Heat load vs. outdoor temperature curve based on average of 22 houses monitored during the first monitoring period, week 26 – 38 ( summer ); week 38 – 47 ( heating season )

6.5 bar

2.2 bar

1.6 bar

65 °C

55 °C

25 °C

dP=0.3bar

elevation = 53 m AMSL

To the main pipe (Lystrup district

heating) 65 °C

Ringgården pumping station

Ringgården low - energy network

Shunt

Return valve (close / half open)

Thermostatic by-pass at each street end

Consumer unit

0 50 100 100 150 200 250 300distance [m]

- 5 0 5 10 15 200

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Week 38 – 47 (y = −85.0 x + 1168 [W], R2 = 0.95)

Week 26 – 38 (108 [W])

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eatin

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[W]

Outdoor temperature [°C]







6.5 bar

2.2 bar

1.6 bar

65 °C

55 °C

25 °C

dP=0.3bar



heating) 65 °C



Shunt



Consumer unit

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- 5 0 5 10 15 200

200

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Week 38 – 47 (y = −85.0 x + 1168 [W], R2 = 0.95)

Week 26 – 38 (108 [W])

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eatin

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ouse

[W]








6.5 bar

2.2 bar

1.6 bar

65 °C

55 °C

25 °C

dP=0.3bar



heating) 65 °C



Shunt



Consumer unit

0 50 100 100 150 200 250 300distance [m]

- 5 0 5 10 15 200

200

400

600

800

1000

1200

1400

1500

Week 38 – 47 (y = −85.0 x + 1168 [W], R2 = 0.95)

Week 26 – 38 (108 [W])

Hea

t loa

d fo

r spa

ce h

eatin

g in

cl.

inst

alla

tion

heat

loss

es p

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ouse

[W]




FIGURE 7: Mean weekly outdoor temperature and indoor temperature in the buildings equipped with room temperature sensors in the living room

FIGURE 8: Average weekly supply, return, shunted supply temperatures and heat load for the 2 year monitoring period

to the storage tank. The real charging flow is thus higher than the design flow. By readjusting this, a lower e(1) can be obtained. The parameter e(1) for the case with IHEU is lower than what is usually used by the designers in similar condi-tions, e.g. 32.3 kW in Denmark. On one hand, this result must be seen in relation to the housing type and residents behav-iours ( mostly senior citizens and young families ). On the other hand, the analysis points at the fact that the dimensioning

although the indoor temperatures during operation were some degrees higher than the design conditions ( 20 °C ), which increased the minimum achievable return temperature from the radiators. Overall, the demonstration project has shown that the concept works, and that is further confirmed by the fact that there were no complaints from residents about the lack and quality of SH / DHW.

Simultaneity factors

In order to define design loads in areas with low - energy buildings, simultaneity curves were developed based on moni-toring data sampled every 4 minutes dur-ing the summer weeks 24 – 38 in the year 2010. Totally 38,000 data sets, or time stamps, were recorded for each of 10 IHEU's and 10 DHSU's. The methodology used to develop the curves is: Data is sorted in the way that the combination of the highest group load E(N) is calcu-lated for each time stamp. To avoid too high simultaneity curves, it is accepted to exceed the suggested design load pr. consumer e(N) in 1 % of the tapping time. This is equal to shortcomings for a time period of 15 minutes if assumed the tap-ping is occurring all 24h / day. In practice the period will be quite shorter than 15 minutes. On the other hand the analysis assumes that the consumers with the highest DHW load by default are placed

at the far end of the DH net. This will sta-tistically not be the case, why this puts the suggested curves to the conservative side. The DHW heat power, e(N), of one consumer was determined to be 4.7 kW for the DHSU case and 24.3 kW for the consumers with IHEU. The e(1) value for the DHSU case is a bit higher than the expected design value of approx. 3 kW. The explanation is the accuracy of the setting of the flow controlled motor valve, which controls the charging flow

26 27 28 29 30 31 32 33 34 35 36 37 38 39 40 41 42 43 44 45 46 47-5

0

5

10

15

20

25

30

Indoor temperature

Outdoor temperature

week

mea

n w

eekl

y te

mpe

ratu

re [ °

C]

T supply shunt, low temperature area T supply DH T return DH Heat load

Week, year

201126, 2010 25, 2012

0

30

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tem

pera

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, wee

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[ °C

]

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, wee

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[ kW

]


Results of a 2 year study in Lystrup, Denmark on a low temperature district heating (DH) system for low energy buildings

Average weekly supply, return, shunted supply temperatures and heat load for the 2 year monitoring period

T supply to low temperature area

‘Delta T’ for low temp DH zone

T supply DH T return DH Heat load

Week, year

201126, 2010 25, 2012

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[kW

]

50 30

38




the end - users and their installations for low - temperature district heating; strate-gies than can be used in a nearby area with 1000 single family houses. The second area is in Høje Taastrup near Copenhagen, where a low - temperature DH system sup-plying 75 houses with floor heating has been built. This area is supplied with an alternative mixing system as described in [ 9 ]: a 3 - pipe shunt arrangement is con-nected to the pumping station supplying mainly DH return water to the low - tem-perature area. When the return tempera-ture is not sufficient, a portion of water from the supply pipe can be added into the mixing shunt. In this case the low - tem-perature network is supplied by water mixed from the supply pipe and the return

Costs

The total investment of the system has been estimated to 346,900 € or approx. 8,460 € per house, see table 5 [ 4 ]. It is also seen from the table, that the DHSU is about 30 % more expensive than the IHEU.

Ongoing R & D project

The investigations continue in an ongoing R & D project where low - temperature dis-trict heating is demonstrated in two exist-ing single - family house neighbourhoods. The first area is in Tilst near Århus, where the low - temperature concept is being tested in a street with 8 houses with radia-tors. The focus is on strategies to prepare

design assumption of e(n) = 1 is leading to an oversize of the DH net. Anyhow, since the smallest available DH branch pipe dimensions are installed, this has no practical impact on the branch pipes. Additional information's for figure 10 can be found in [ 10 ].

Distribution heat losses and pump electricity consumption

For a Danish reference year DH demand, heat loss in the distribution network and the annual electricity use of the pump were calculated based on duration curves divided in 8 representative inter-vals combined with load vs. temperature curves derived for the first monitoring period of 2010 ( as figure 6 ). In addition, a full year ( 2011 ) of measurements is available, see table 4.The heat loss of the reference year for the entire network is in line with the expected heat loss calculated in the design phase and comparable with the present share of the heat loss in the existing city - wide dis-tribution networks in Denmark. However, the measured heat loss for the full year 2011 is about 11 % higher, which can be explained by a distance of unintended non - insulated pipes before the DH meter in each dwelling with IHEU. Considering these pipes insulated, the distribution heat loss for a network with 11 IHEU's is only slightly above the corresponding distribu-tion heat losses for the 11 DHSU's. The total distribution heat loss in the low-tempera-ture network are approx. ¼ of the esti-mated heat loss in the case of conventional medium - temperature network ( single pipes, series 1, 80 / 40 °C, 6 bar system, 1 m / s flow velocity ). The electricity use for pump-ing was estimated to be 2,600 kWh / yr, equivalent to 9 kWhel /MWhth. This is com-parable with the electricity demand for pumping purposes in existing well - estab-lished systems [ 4 ]. According to the design method, it was expected to measure a larger pumping demand; the lower elec-tricity use for the pump is explained in practice by the fact that the pressure levels in the network were still well below the lim-its set. This points that there is room for optimizing the network design method even more, so that the heat loss can be sig-nificantly decreased, at expenses of an additional, but less significant from the overall primary energy point of view, pumping demand. Hydraulic limitations and noise must not be forgotten, though.

year

DK ref ( calc. )

2011 ( meas. )

total heat delivered to LTDH network MWh 287.2 273.9

heat demand MWh 238.1 219.4

distribution heat lossMWh 49.1 54.5

% 17.1 19.9

heat power, yearly avg. kW - 31.3

supply temperature, DH °C - 67.4

supply temperature, LTDH °C 55 52.7

return temperature DH °C 30 34.1

electricity use, pumping station kWh 2600 2566

TABLE 4: Key data of network operation

costs ( 2010 )

item [ € / m ] [ € / unit ] total [ € ]

pipes* 120 65,000

pipe fittings* 32 17,000

pipe laying** 131 100,500

DHSU substation* 3,700 41,000

IHEU substation* 2,600 78,000

substation installation** 1,000 41,000

pump + frequency controller* 2,400 + 2,000 4,400

total cost 346,900

cost per house 8,460

TABLE 5: investment costs



In the case considered, the distribution heat loss for the area with DHSU's are slightly lower than in the area with IHEU's. The sum of the distribution heat loss and the standby heat loss from the substation is on the other hand larger in the DHSU case than in the case with IHEUs, because the additional heat loss due to the stor-age tanks more than counteracts the reduction of the distribution heat loss. However, in areas with hydraulic limita-tions, such as outer urban areas, DHSUs offer in turn some advantages, thanks to the lower peak pressure / load require-ments. Moreover, the smallest media pipe diameters of the house connection pipes in the market have still a valuable water flow overcapacity and this suggest that smaller volume of the storage tank can be chosen, in case of DHSU, and this would reduce the substation heat loss, space occupation and costs somehow. The conclusion is that within the tested substations, the IHEU is a better solution in regards to energy performance, instal-lation costs and space requirements. Any-how there is no superior substation con-cept for all purposes, but the system should be chosen taking into account the specific characteristics of the site and of the demand.

Aknowledgement

The projects have received grants from the Danish EUDP - program making it possible to develop and demonstrate the low -tem-perature concept.

pipe of the main district heating network. This solution can be installed in an existing district heating network at a location hav-ing a sufficient flow in the return pipe. In addition the monitoring continues in Lys-trup with further analysis of substations and distribution network.

Conclusions

The demonstration project of a low - tem-perature DH network for low - energy buildings has shown that the concept works. The results show that it is possible to supply the customers with a supply temperature of approx. 50 °C and satisfy both the SH requirements and the safe provision of DHW. This fact is confirmed by the fact that there were no complaints from residents about the lack of SH or DHW. The energy efficiency target was met, being the distribution heat loss equal to 17 % of the total heat production for the Danish reference year. Even better real - life performance is expected when unintended non - insulated pipes are get-ting insulated.In DH networks of this kind, serving low heat density areas with no possibilities for future expansion, the design should envisage the exploitation of the maxi-mum pressure that can be withstood by the media pipes. The network design method can thus be optimized, so that the distribution heat loss can decrease even further, at expenses of an additional, but less significant, pumping demand.The analysis points at the fact that the dimensioning of DH systems need a bet-ter basis for simultaneity factors and that a greater consideration must be given to the operation of the SH and DHW instal-lations, for the calculation of the optimal size of the heat distribution system.The results demonstrate that it is possible to guarantee an energy - efficient opera-tion, but it is very important to obtain proper functioning of each substation, otherwise unacceptable return tempera-tures result.



In the case considered, the distribution heat loss for the area with DHSU's are slightly lower than in the area with IHEU's. The sum of the distribution heat loss and the standby heat loss from the substation is on the other hand larger in the DHSU case than in the case with IHEUs, because the additional heat loss due to the stor-age tanks more than counteracts the reduction of the distribution heat loss. However, in areas with hydraulic limita-tions, such as outer urban areas, DHSUs offer in turn some advantages, thanks to the lower peak pressure / load require-ments. Moreover, the smallest media pipe diameters of the house connection pipes in the market have still a valuable water flow overcapacity and this suggest that smaller volume of the storage tank can be chosen, in case of DHSU, and this would reduce the substation heat loss, space occupation and costs somehow. The conclusion is that within the tested substations, the IHEU is a better solution in regards to energy performance, instal-lation costs and space requirements. Any-how there is no superior substation con-cept for all purposes, but the system should be chosen taking into account the specific characteristics of the site and of the demand.

Aknowledgement

The projects have received grants from the Danish EUDP - program making it possible to develop and demonstrate the low - tem-perature concept.

pipe of the main district heating network. This solution can be installed in an existing district heating network at a location hav-ing a sufficient flow in the return pipe. In addition the monitoring continues in Lys-trup with further analysis of substations and distribution network.

Conclusions

The demonstration project of a low - tem-perature DH network for low - energy buildings has shown that the concept works. The results show that it is possible to supply the customers with a supply temperature of approx. 50 °C and satisfy both the SH requirements and the safe provision of DHW. This fact is confirmed by the fact that there were no complaints from residents about the lack of SH or DHW. The energy efficiency target was met, being the distribution heat loss equal to 17 % of the total heat production for the Danish reference year. Even better real - life performance is expected when unintended non - insulated pipes are get-ting insulated.In DH networks of this kind, serving low heat density areas with no possibilities for future expansion, the design should envisage the exploitation of the maxi-mum pressure that can be withstood by the media pipes. The network design method can thus be optimized, so that the distribution heat loss can decrease even further, at expenses of an additional, but less significant, pumping demand.The analysis points at the fact that the dimensioning of DH systems need a bet-ter basis for simultaneity factors and that a greater consideration must be given to the operation of the SH and DHW instal-lations, for the calculation of the optimal size of the heat distribution system.The results demonstrate that it is possible to guarantee an energy - efficient opera-tion, but it is very important to obtain proper functioning of each substation, otherwise unacceptable return tempera-tures result.


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