Oversized Heat Exchangers

The effect of over sizing a heat exchangerThe previous calculations (Tutorial1 3.2) assumed that the heat exchanger had been sized on the perfect heating area to meet the specification. This would mean that the heat exchanger was exactly sized for the duty.

This is highly unlikely in practice as the designer or specifier will usually add other factors, including those for fouling and uncertainty of maximum operating loads. It is also unlikely that manufacturers can supply heat exchangers to match a specification exactly. As undersized heat exchangers are impractical they are usually bought oversized.

The operating conditions laid down in Example 13.2.1, Part 'C', have been reconsidered in Example 13.3.1 by adding 15% to the required heating area to account for contingencies.

Required heating area is calculated to be 1.09 m2 (Example 13.2.1, Part 'C') therefore the specified heating area for Example 13.3.1 is to be 1.09 + 15% = 1.254 m2.

The minimum size that the manufacturer can supply has a heating area of 1.31 m 2, representing an actual heating area of some 20% above that required. A larger heating area requires less steam pressure for the same heat transfer rate, and because of this the steam pressure in an oversized heat exchanger will be lower for the same heat load.

As the steam pressure is less, the steam temperature is less, and the heat exchanger LMTD (Logarithmic Mean Temperature Difference) will also be less.

To determine the steam temperature for the design condition, it is first necessary to find the new LMTD (DTLM) for the larger heating area (see Example 13.3.1).

Example 13.3.1The ΔT LM can be found by re-arranging Equation 13.2.1 to give Equation 13.3.1

Equation 13.2.1

Equation 13.3.1

Where:

ΔT M = Mean temperature difference. Note: ΔT M may be either ΔT LM (LMTD) or ΔT AM (AMTD)= Mean heat transfer rate (W)

U = Heat transfer coefficient (W/m2°C)A = Heating area (m2)

From Example 13.2.2, at full-load:

The secondary inlet temperature (T 1= 10°C

The secondary outlet temperature(T 2) = 60°C

          The new steam design temperature can now be determined using Equation 2.5.5:

Equation 2.5.5

Where:

ΔT LM = 95.95 CT 1 = 10°CT 2 = 60°CT S = Steam temperature °C

By taking antilogs of both sides of the equation . . .

his temperature corresponds to a steam pressure of 1.95 bar g. When the heat exchanger was perfectly sized in Tutorial 13.2, the steam pressure was 4 bar g. In this example, with a heat exchanger 20% oversized, the steam pressure is 51% less.

Now that the steam pressure has been predicted at the full-load condition, it is possible to calculate the steam flow at full-load.

By using Equation 2.8.1 find the steam flowrate at the full heat load of 314.25 kW. At 1.95 bar g, steam tables state that the enthalpy of evaporation is 2164.6 kJ/kg.

Equation 2.8.1

The steam flow was 536.6 kg/h in the perfectly sized heat exchanger (Example 13.2.1), so it can be seen that there is a slight drop (2.5%) in mass flowrate. This is due to the steam having a slightly larger enthalpy of evaporation in the larger heat exchanger due to its lower pressure.

Determine the TDC for the larger heat exchanger

Now that the steam temperature has been determined for the oversized heat exchanger (using the LMTD equation [Equation 2.5.5]), it is now possible to find its TDC, using Equation 13.2.2.

Equation 13.2.2

Where:

TDC = Temperature Design ConstantT S = 133.1°CT 1 = 10°CT 2 = 60°C

At the minimum heat load:

When the heat exchanger was perfectly sized in Example 13.2.1 the steam temperature was 115.2°C at the minimum heat load of 188.5 kW.

ecause the oversized heat exchanger in this example is about 20% larger, the steam temperature will also be less at the minimum heat load. The minimum heat load remains the same as in Example 13.2.1 and occurs when the secondary inlet temperature rises to 30°C.

From Equation 13.2.3:

Equation 13.2.3

Where:

T S = Steam temperature °C T 1 = Secondary fluid inlet temperature °C = 30°CT 2 = Secondary fluid outlet temperature °C = 60°CTDC = Temperature Design Constant = 1.684

Comparing the two heat exchangers at minimum load, the steam temperature has dropped from 115.2°C in the perfectly sized heat exchanger to 103.8°C in the oversized heat exchanger.

From steam tables, this steam temperature corresponds with a steam pressure of about 0.15 bar g, and hfg = 2 247 kJ/kg. The steam pressure in the perfectly sized exchanger (at 115.2°C) was 0.7 bar g.

By using Equation 2.8.1, it is possible to find the steam flow at the minimum heat load of 188.5 kW.

Equation 2.8.1

The minimum steam flow was 306 kg/h in the perfectly sized heat exchanger (Example 3.2.1), so it can be seen that there is a marginal drop in mass flow in the oversized heat exchanger at the minimum heat load. This is due to the steam having a slightly larger enthalpy of evaporation in the larger heat exchanger due to its lower pressure.

The steam pressure, the steam trap, and effective condensate removal

As the steam gives up its heat across the heat transfer surface to the secondary fluid, it condenses in the steam space. Condensate passes out through the outlet of the heat exchanger, and through a steam trap, which traps the steam in the steam space whilst allowing the condensate to be freely discharged.

If the heat exchanger has not been specifically designed to operate with condensate flooding the steam space, the steam pressure needs careful consideration to ensure the heat exchanger is properly drained of condensate. Any waterlogging of the steam space will reduce the effective heating surface area, and the heat transfer requirement may be satisfied only if the exchanger is sufficiently (perhaps accidentally) oversized.

The capacity of the steam trap will depend upon its type, its orifice size and the differential pressure across it. Differential pressure provides the energy to push the condensate through the trap, and is the difference between the steam pressure in the heat exchanger, and the backpressure exerted on the outlet of the trap by the condensate system.

If the steam trap drains by gravity via a properly sized pipe to a vented condensate receiver or an open end, the

backpressure should be very near atmospheric. Under these conditions, the differential pressure on a sizing chart can simply be read as the gauge pressure in the heat exchanger.

If, however, there is a lift after the trap (a rise in the trap discharge line), or the trap discharge line is undersized, or this line is pressurised for any other reason, the backpressure may, at times, be greater than the pressure in the steam space. When this is so, the differential pressure across the trap is reversed and is deemed to be a 'negative differential pressure'. The trap capacity is now zero.

As can be seen in the above calculations, the steam pressure in any heat exchanger is governed by its size and the secondary conditions. As the capacity of the steam trap depends on the differential pressure, it follows that changes in the steam pressure and backpressure affect the capacity of the steam trap at all times. As the differential pressure reduces, the capacity of the steam trap will fall. Provided the differential pressure is positive and the steam trap is selected and sized with this in mind, waterlogging and its associated problems will not occur.

Sizing the steam trap for the oversized heat exchangerThe conditions that need consideration are:

Full-load : 523 kg/h at 1.95 bar g in the steam space

Minimum load: 302 kg/h at 0.15 bar g in the steam space

Backpressure: Atmospheric pressure (0 bar g)

Figure 13.3.1 Static head and vacuum breaker method of dealing with stall

Consider, on the float trap capacity chart Figure 13.3.2, a DN25 (1") FT14-4.5 ball float steam trap. It can be seen that it will pass 850 kg/h at a differential pressure of 1.95 bar. It may also be seen that at a differential pressure of 0.15 bar it will pass about 370 kg/h. In this example, consider the trap fitted to the oversized heat exchanger and draining by gravity to a vented condensate receiver, as depicted in Figure 13.3.1.

To ensure proper drainage, the steam trap has to be able to cope with all loads between the full-load and minimum load conditions.

As the condensate backpressure is atmospheric in this example, the minimum steam space pressure of 0.15 bar g is always higher than the backpressure. It can be seen from the capacity chart (Figure 13.3.2) that the trap has enough capacity at the minimum and maximum loads, so the DN25 (1") FT14-4.5 ball float steam trap is big enough.

If, however, in this example, the backpressure were higher than the minimum steam pressure of 0.15 bar g, the system would stall somewhere within the normal operating range. (This would only require a lift of just more than 1.5 metres after the trap to cause this). Accordingly, the trap would have to be selected and sized depending upon the amount of backpressure. With larger amounts of backpressure it may be necessary to fit a pump-trap.

Advice on how to select the correct trap for a heat exchanger is given in Tutorial 13.4.

Fig. 13.3.2 FT14 ball float steam trap capacity chart showing data for Example 13.3.1

The Heat Load, Heat Exchanger and Steam Load Relationship

Calculations for heat exchange applications including the design loads and the steam pressure/flowrate requirements.

Saturated steam is used to provide primary heat to a process fluid in a heat exchanger. The term heat exchanger is used to describe all types of equipment where heat transfer is promoted from one fluid to another. For convenience, this broad definition will be applied to the term heat exchanger. While shell and tube heat exchangers and plate heat exchangers will be principally referred to, stall may also be relevant to applications including air heater batteries, submerged tank coils, jacketed vessels and storage calorifiers.

Temperature controlled applicationsIn a temperature control application, the inlet temperature of the secondary fluid to the heat exchanger may change with time. This means that in order to maintain a consistent secondary fluid outlet temperature, the heat supplied to the heat exchanger must also vary. This can be achieved by using a control valve on the inlet to the primary side of the heat exchanger, as shown in Figure 13.2.1.

Fig. 13.2.1 Typical temperature control of a steam/water shell and tube heat exchanger

A control valve is used to vary the flowrate and pressure of the steam so that the heat input to the heat exchanger can be controlled. Modulating the position of the control valve then controls the outlet temperature of the secondary fluid. A sensor on the secondary fluid outlet monitors its temperature, and provides a signal for the controller. The controller compares the actual temperature with the set temperature and, as a result, signals the actuator to adjust the position of the control valve.

For a constant heating area and heat transfer coefficient, the rate at which heat is transferred from the steam to the secondary fluid for a particular heat exchanger is determined by the mean temperature difference between the two fluids. A larger difference in mean temperatures will create a large heat transfer rate and vice versa. On partially closing the control valve, the steam pressure and the temperature difference fall. Conversely, if the control valve is opened so that the steam mass flow and hence pressure in the heat exchanger rise, the mean temperature difference between the two fluids increases.

Altering the steam pressure will also slightly affect the amount of heat energy available in the condensing steam as the enthalpy of evaporation actually falls with increasing pressure. This means that the latent heat available per kg of steam reduces as the steam pressure increases. If steam flow accuracy is required, this must be accounted for.

Example 13.2.1

A manufacturer is to design a heat exchanger in which the specification calls for steam at 4 bar g to heat secondary water from 10°C to 60°C. The water flow is to be constant at all loads at 1.5 L/s. It is assumed that 1 litre of water has a mass of 1 kg, so the mass flowrate = 1.5 L/s x 1 kg/L = 1.5 kg/s. The manufacturer uses a heat transfer coefficient 'U' for the heat exchanger of 2500 W/m2°C.Take the specific heat of water as 4.19 kJ/kg°C.

Determine:

(A) The design heat load.

(B) The corresponding steam flowrate.

(C) The minimum heating area required.

A. Also, if the customer's minimum heat load occurs when the inlet water temperature rises to 30°C, determine:

(D) The minimum heat load.

(E) The corresponding steam pressure in the heat exchanger.

(F) The corresponding steam flowrate.

Calculations:

(A) Find the design heat load using the heat transfer flowrate equation (Equation 2.6.5):

Equation 2.6.5

Where:= Mean heat transfer rate (kW)

= Mean secondary fluid flowrate (kg)

cp = Specific heat capacity of the secondary fluid (kJ/kg K) or (kJ/kg°C)

T = Temperature rise of the secondary fluid (K or °C)

(B) Find the corresponding steam flowrate at 4 bar g, saturation temperature (T s) is 152°C, and hfg = 2108.1 kJ/kg (from steam tables). Calculate the required steam flow at the design condition using Equation 2.8.1:

Equation 2.8.1

(C) Find the minimum heating area to meet the requirement using Equation 2.5.5. Note; the manufacturer uses the Logarithmic Mean Temperature Difference (ΔT LM) to calculate the minimum amount of heating area to satisfy the design rating:

Equation 2.5.5

Where:

ΔT LM = Logarithmic Mean Temperature Difference (LMTD)T s = Steam temperature (°C)T 1 = Secondary fluid in temperature (°C)T 2 = Secondary fluid out temperature (°C)ln = The mathematical function known as 'natural logarithm'

By re-arranging the general heat transfer equation (Equation 2.5.3: = U x A x T) Equation 13.2.1 can be formulated, where ΔT can be represented by the mean value ΔT M.

Equation 13.2.1

Where:

A = Heating area (m2)= Mean heat transfer rate (W)

U = Heat transfer coefficient (W/m2°C)T M = Mean Temperature Difference. Note: ΔT M may be either ΔT LM (LMTD) or ΔT AM (AMTD).

For the purpose of this example it will be assumed that the heat exchanger is designed to have exactly this area of 1.09 m2.

(D) Find the minimum heat load, when the inlet water temperature is 30°C, using the heat transfer flowrate equation (Equation 2.6.5) as used in Part 'A' of these calculations:

Equation 2.6.5

To calculate the corresponding steam flowrate, it is first necessary to determine the steam temperature at the minimum load condition.

It is possible to use the T LM design figures to accurately predict the steam temperature for any load condition, but this requires the use of logarithmic calculations. However, once the exchanger size is fixed and the design temperatures are known, it is much easier to predict operating temperatures using what could be termed a heat exchanger Temperature Design Constant (TDC). The TDC method does not require logarithmic calculations. Please note: TDC cannot be used on those applications where the secondary flowrate varies or where control is achieved by varying the condensate level in the steam space.

Note: When sizing a heat exchanger it is normal for heat exchanger manufacturers to use the T LM method. Once sized, by knowing the heating area and the full-load operating temperatures, TDC can be used to accurately predict all operating temperatures resulting from changes in load, as can be seen in the following text.

Operating temperatures can also be predicted graphically by using what is termed a 'Stall Chart'. This method is discussed in Tutorials 13.5, 13.6, and 13.7.

Temperature Design Constant (TDC)

For any type of steam-heated exchanger with the secondary liquid flowing at a constant rate, TDC can be calculated from the test figures quoted by the manufacturer for full-load. If these data sets are not available and the heat exchanger is already installed in service, TDC can be calculated by observing the steam pressure (and finding the steam temperature from steam tables) and the corresponding secondary inlet and outlet temperatures at any load.

TDC is the ratio of the steam to water temperatures at the inlet and outlet; and is shown in Equation 13.2.2.

Equation 13.2.2

Where:

TDC = Temperature Design ConstantT s = Steam temperatureT 1 = Secondary fluid inlet temperatureT 2 = Secondary fluid outlet temperature

In Example 13.2.1 at full-load conditions:

The TDC equation can be transposed to find any one variable as long as the other three variables are known. The following equations are derived from the TDC equation (Equation 13.2.2).

To find the steam temperature at any load use Equation 13.2.3:

Equation 13.2.3

To find the secondary fluid inlet temperature at any load use Equation 13.2.4:

Equation 13.2.4

To find the secondary fluid outlet temperature at any load use Equation 13.2.5:

Equation 13.2.5

For any heat exchanger with a constant secondary flowrate, the operating steam temperature can be calculated for any combination of inlet temperature and outlet temperature.

In Example 13.2.1 the secondary outlet temperature remains at 60°C, and minimum load occurs when the inlet temperature is 30°C. What is the steam temperature at minimum load?

Inlet temperature = 30°C

Outlet temperature = 60°C

Using Equation 13.2.3:

(E) Find the corresponding heat exchanger steam pressure and enthalpy at minimum load

From steam tables:

A steam temperature of 115.2°C corresponds with a steam pressure of 0.7 bar g.

The specific enthalpy of evaporation at 0.7 bar g (hfg) = 2 215 kJ/kg

(F) Find the steam flowrate at minimum load:

From (D) the minimum heat load is 188.5 kW.

From (E) the hfg is 2 215 kJ/kg.

Using Equation 2.8.1:

Equation 2.8.1

Condensate removal in Heat ExchangersThis Block discusses the removal of condensate from heat exchange equipment supplied by saturated steam and fitted with:

A temperature control valve on the steam line to the heat exchanger.

A steam trapping device on the condensate line from the heat exchanger.

The primary side of the heat exchanger will be referred to as the 'steam space', and the steam trapping device will be referred to as the 'trap'. The 'trap' can be a 'steam trap', a 'pump trap', or a 'steam trap and pump' fitted in combination.

On these installations, a control sensor monitors the temperature of the outgoing heated fluid in the secondary circuit. The control valve endeavours to maintain a temperature determined by the controller, regardless of variations in heat load. The valve achieves this by opening or closing to alter the flowrate of steam, thereby varying the steam space pressure.

The discharge from the steam trap may be subject to a lift and/or pressure in the condensate line, or may fall to an open end where it is subjected only to atmospheric pressure. This Block will refer to condensate pressure as 'backpressure'.

The heat exchange equipment can be almost anything that meets the above criteria. Examples include:

Shell and tube heat exchangers.

Plate heat exchangers.

Air heating coils or batteries in ductwork.

Pipe runs or pipe coils in process equipment, tanks, vats etc.

For brevity, this Block will refer to all such devices as 'heat exchangers' or 'heaters', and the passage of fluid being heated by the heat exchanger will be referred to as passing through the 'secondary' side of the heat exchanger.

The performance of steam heat exchangers is often reduced due to condensate flooding the steam space and waterlogging. The two main causes of waterlogging are:

Fitting the wrong type of trap.

Stall.

Important noteSome systems aim to achieve control of temperature by positively encouraging partial flooding of the steam space of the heat exchanger. In these cases, the modulating action of the control valve at the condensate outlet varies the condensate level in the steam space. This changes the area of heating surface exposed to steam, and the effect is to change the heat transfer rate so as to control the secondary outlet temperature.

With systems of this type, it is important that the heat exchangers be designed and manufactured specifically to withstand the effects of flooding. Where this is not done, the presence of condensate in the heat exchanger will have an adverse effect on operating performance and will reduce service life.

This method of control can have certain benefits if the system is designed correctly. One is that the condensate sub-cools in the heat exchanger before it is discharged. This can considerably reduce the amount of flash steam in the condensate pipework, which may improve the performance of the condensate system and also reduce heat losses.

The main operational disadvantage is that systems of this type are slow to respond to variations in heat load.

What is meant by stall?Stall is the reduction or the cessation of condensate flow from the heat exchanger, and occurs when the pressure in the heat exchanger is equal to, or less than, the total backpressure imposed on the steam trap.

Lower than expected pressure in a heat exchanger may occur as a result of any of the following circumstances:

The secondary fluid inlet temperature rising as a result of a falling heat load.

The secondary fluid flowrate falling as a result of a falling heat load.

The secondary fluid outlet temperature falling due to a lowering of the set point.

As the control valve reduces the steam pressure to meet a falling heat load, the lack of differential pressure across the steam trap causes condensate to waterlog the steam space, as shown in Figure 13.1.1.

Fig. 13.1.1 An air heater battery suffering the effects of stall

Due to applied safety factors and because heat exchangers are sold in pre-determined sizes, they often have more heating area than required. This has the effect of increasing the heat transfer capability of the exchanger above that required. It also means that the operating steam pressure will be lower than in a comparable heat exchanger perfectly sized for the same duty. The result is that less steam pressure is available to push out the condensate than may be expected. The steam pressure in the heat exchanger is important because it influences the stall condition, which in turn affects trap selection.

Before any trap selection and sizing can take place, it is necessary to determine whether or not stall will occur, and if it does, to what degree. If this is not done, it is likely that the heat exchanger will suffer from waterlogging for some or all of its operating life. This, when it occurs, may not be immediately recognised by the observer or operator, as operating performance might not be reduced in an oversized heat exchanger. However, waterlogging can have severe financial consequences, short and long term, unless the heat exchanger is designed to operate this way.

Short-term problemsConsider an oversized heater battery operating as a frost coil and fitted with the wrong type (or size) of trap, as in Figure 13.1.1.

In this example, the frost coil is preheating chilled air before it passes on to the main heater battery. Though the frost coil is fulfilling its thermal expectations (because it is oversized for the duty), it will do so with the bottom half of its coils waterlogged. Incoming cold air approaching 0°C (typically flowing at 3 m/s) passing over the coils can easily cause the water in them to freeze. This results in having to repair or replace the heater battery, either causing inconvenience or unexpected outlay.

Waterlogging and freezing will not arise if the application is correctly designed.

Long-term problems Traps that are undersized will sometimes show no immediate adverse effects on heater performance if the heater is oversized.

Ironically, the wrong type of trap fitted to a heat exchanger can often exaggerate a superficial improvement elsewhere in the condensate system. For instance, a thermostatic or fixed orifice fitted to any heat exchanger will hold back condensate so that it sub-cools below the steam saturation temperature. This will have the effect of reducing flash steam from any natural outlet such as a condensate receiver vent. The casual observer can interpret this as a way to

save energy and can easily be tempted to fit these devices. Unfortunately, the situation is not as straightforward as it seems. The reality is that holding back condensate until it sub-cools implies waterlogging to some degree. Condensate that continually floods the steam space will cause corrosion with costly results. The service life of the heat exchanger is reduced, and the overall lifetime costs of the installation will increase.

The effects suffered by a waterlogged heat exchanger depend upon the circumstances of the particular installation.

The symptoms and effects of stall are itemised later in this Tutorial.

How does stall occur?To understand stall it is necessary to appreciate that saturated steam is a condensing vapour, which gives up its heat as it condenses to water. This condensation always occurs at a constant temperature when the pressure in the steam space remains constant.

For example, saturated steam at atmospheric pressure has a temperature of 100°C and will also condense back into water at 100°C, whereas at a gauge pressure of 1 bar, saturated steam has a temperature of 120°C and will condense back into water at 120°C. Steam can also exist inside heat exchangers at below atmospheric pressure i.e. steam at 0.5 bar below atmospheric pressure has a temperature of about 82°C, and will also condense back to water at 82°C. The pressure and temperature relationship of saturated steam is entirely predictable and is documented in steam tables.

Basic heat exchanger theory states that the higher the steam temperature above that of the secondary fluid being heated, the greater the potential heat transfer rate. To vary the transfer of heat from condensing steam, the temperature (and thus the pressure) of the steam in the steam space is varied.

For example, if a heat exchanger uses steam at 160°C at maximum load, and the load is reduced by 50%, steam at a lower temperature is required. To achieve this, the steam pressure must be reduced, and, in many cases, becomes less than the backpressure.

Example:A heat exchanger running at full-load uses saturated steam at 1 bar g (120°C) to heat water from 40°C to 60°C. Full-load therefore occurs when the water temperature rises by 20°C, and the mean water temperature is:

The difference between the steam temperature and the mean water temperature is termed the Arithmetic Mean Temperature Difference or AMTD, and the heat transfer rate is proportional to this. The full-load AMTD in this example is 120°C - 50°C = 70°C.

Consider the situation where the process load falls to 2/3 load.

At full-load, the water temperature rise is 20°C.

If the load falls to 2/3 full-load, and the outlet water temperature remains constant at 60°C, this means that the temperature rise must be 2/3 of 20°C

Therefore:At 2/3 load, temperature rise = 2/3 of 20°C = 13.3°Cand the inlet temperature = 60°C - 13.3°C = 46.7°C

Consequently at 2/3 load, the return water temperature will have risen to 46.7°C, and so the mean water temperature is now:

At 2/3 load, the heat transfer needed will be 2/3 of that at full-load, and equally the AMTD will be 2/3 of that at full-load, i.e.

It follows that the steam temperature at 2/3 load has to be the mean water temperature at 2/3 load plus the AMTD at 2/3 load, i.e.

As the temperature of saturated steam at atmospheric pressure is 100°C, this means that the pressure in the steam space is now atmospheric. Consequently, there is no steam pressure available in the steam space to push the condensate through a steam trap. Even if the condensate line fell to an open-ended steam trap, the condensate might not drain out of the exchanger. The condensate will 'back-up' the drain line and waterlog the heat exchanger unless proper precautions are taken.

If condensate backs up into the exchanger, the surface area available to condense steam is reduced, the heat flow drops and the temperature of the outgoing heated water begins to fall. When the temperature sensor detects this, the controller opens the control valve a little more and the inflow of steam increases. This raises the pressure in the steam space above atmospheric (in this case) and soon becomes high enough to push condensate through the trap. The condensate level falls, but now the steam space pressure is higher than the atmospheric pressure needed to just heat the water to 60°C. The water temperature then climbs. When the sensor detects this, the controller closes down the control valve. The steam space pressure falls to atmospheric - and the flooding begins again.

The result is a continual cycling of the water temperature above and below 60°C. If the secondary medium were other than water this could, in many cases, affect its quality.

What are the symptoms and effects of stall?One or more of the following symptoms may be evident:

In summary:

1. Cold or cool steam trap. 2. Hunting control valve.

3. Fluctuating outlet temperature.

4. Stratified heater temperatures.

5. Waterhammer.

6. Reduced heat output.

7. Reduced product quality.

8. Corroding heat exchangers.

9. Leaking heat exchangers.

10. Failing heat exchangers.

In detail:

The steam trap goes cold, or is noticeably cooler than the temperature of the steam pipe inlet to the heat exchanger.

The control valve is prone to 'hunting', i.e. it cycles regularly somewhere between its open and closed positions.

The temperature of the secondary fluid flowing from the heat exchanger is less accurate than is expected or required.

There is stratification of temperature on the output side of the heat exchanger. This will be more apparent on heater batteries and unit heaters.

For example, it is almost certain to be detectable on the air heater battery depicted in Figure 13.1.1. The design is such that the face of the heat exchanger surface is usually accessible, often via an access panel or door in the side of the ducting. If stall is happening, the top of the battery closest to the steam inlet will be very hot, whereas lower down, it will be much cooler or even cold, and the trap will be cool or cold. The temperature of the air flowing through the top of the battery will be noticeably higher than that flowing through the bottom.

The heat exchanger makes crackling, banging or thumping noises either continuously or intermittently. Sometimes these noises are associated with severe waterhammer that can cause physical damage to the heat exchanger and any equipment fitted to it. The hot steam condensing into the waterlogged condensate causes the waterhammer and resulting noises, especially when the waterlogging level varies with changes in load.

In process applications, the result of one or more of the above symptoms may be poor or unreliable product quality.

Increased corrosion. The waterlogged condensate cools to temperatures much lower than the steam temperature at the inlet to the steam space. Carbon dioxide and oxygen dissolve much more readily into cooler water.

Carbon dioxide is a common by-product of incorrect boiler water treatment and is carried over into the heat exchanger with the steam. When it dissolves into water it forms carbonic acid, which causes corrosion.

Oxygen is present in raw water, and if not completely removed by the water treatment process, it too will get carried over with the steam. Its presence in water, especially cool water in which it will readily dissolve, also aggravates corrosion.

Corrosion rates are greatly accelerated when both gases are present.

The degree of corrosion will depend upon the heat exchanger material. Copper, carbon steel, and stainless steel will each be affected differently.

Mechanical stress.

The hot steam in the top of the steam space will cause the heat exchanger to expand there, while the cool water in the bottom of the steam space has the reverse effect. This uneven expansion/contraction can cause mechanical stress to the heat exchanger structure, notably to the soldered, brazed, welded or expanded joints in 'plate' and 'shell and tube' heat exchangers, and air heater batteries. The most common result is leakage of steam to the surroundings in the former, or into the secondary airflow in the latter. The stress tends to be worse if the waterlogging level continually varies, especially if it varies quickly. The level of waterlogging will vary as the load changes, and as a result; the control valve and steam trap will struggle to achieve stable control.

It should be said that a properly engineered plate heat exchanger with gasket joints suitably designed for steam will be very resilient to such stress.

The ultimate effect of stall is increased maintenance and shorter service life of the heat exchanger and associated equipment. This increases overall running costs.

Do all heat exchangers suffer from stall?No. The conditions may be such that there will always be sufficient positive pressure upstream of the steam trap to clear the condensate so stall cannot occur.

As a general rule, the higher the secondary temperature above 100°C, and the more stable the running load, (especially if near to the maximum output of the heat exchanger), the less likely for stall to occur. However, each application is unique and will require individual consideration. The only ways to determine the dynamics of the installation are to either plot the application temperatures on a chart or to perform a mathematical calculation. This is explained in Tutorial 13.2, 'Condensate Removal from Heat Exchangers'.

Some applications can appear to operate with partial waterlogging, and show little effect of waterhammer. These tend

to be steady load applications, or where the load changes only slightly and very slowly, and/or applications that employ very robust heat exchange equipment.

One such example would be large bore corrosion resistant heating coils inside tanks correctly arranged to have a positive fall towards the trapping points.

Even in applications of this type, if the installation is designed or corrected to eliminate stall, improved operation, improved reliability, and reduced lifetime costs are virtually guaranteed.

If stall conditions are inevitable, potential problems can be overcome by designing the installation around one of three basic solutions:

1. Ensure the steam pressure in the steam space can never drop below atmospheric pressure, and that the condensate can drain by gravity to and from a ball float steam trap.

2. Accept that the pressure in the steam space may be less than the backpressure, and provide an alternative means of removing condensate, by installing a pump-trap.

3. Ensure the pressure in the steam space is stable and higher than the backpressure. This will entail having the temperature control system on the secondary side of the system.

Taking these three options in turn:

1. Installations that ensure the conditions in the steam space can never drop below atmospheric pressure, and that the condensate can drain by gravity to and from a steam trap:

1a) Condensate removal by vacuum breaker method (see Figure 13.8.1)The steam trap cannot be subject to any backpressure higher than atmospheric, and must drain condensate either to an open end (which may be wasteful), or to a nearby vented receiver and pump, enabling the energy contained in the condensate to be reclaimed.

There are two criteria that must be satisfied:

A vacuum breaker must be fitted to the steam inlet to the heat exchanger after the control valve.

The trap must be installed at a discreet distance below the heat exchanger outlet such that sufficient static head is created to pass the requisite amount of condensate when stall occurs. A distance of between 0.5 to 1 m is usually sufficient; however, smaller distances can be accommodated with larger traps, if less head is available.

Fig. 13.8.1 Static head and vacuum breaker method of dealing with stall

1b) Auxiliary drain trap method (see Figure 13.8.2)A standard float trap set is installed with condensate returning to a condensate system, which is either pressurised and/or elevated above the trap. An auxiliary float trap may be fitted, discharging condensate via an open end to drain.

When there is sufficient steam pressure to overcome the backpressure, the main float trap will function, but when stall occurs, condensate will back-up and drain through the auxiliary float trap thus preventing condensate flooding back into the heat exchanger.

As this condensate will drain to waste, this method should only be used if stall occurs infrequently. The auxiliary trap should be sized on static head to pass the stall load as in method 1a, and the 'main' trap should be the same size, but fitted at least 150 mm below the auxiliary take-off tee-piece.

Apart from the obvious disadvantage of energy loss, this method also requires available head between the trap inlets and the heat exchanger outlet.

Fig. 13.8.2 Auxiliary drain method of dealing with stall

2. Installations which allow the steam pressure in the steam space to drop below the backpressure, but where the condensate can drain by gravity to a pump-trap arrangement:

2a) A pump and float trap installed in combination (see Figure 13.8.3)This method uses a pump and float trap installed in combination. It is better suited to heat exchangers with nominal heating capacities in excess of 1.5 MW (nominally 2500 kg/h of steam).

The steam pressure changes relative to changes in heat load. At high loads the steam pressure will be higher than the backpressure, but at low loads it will be lower.

The pump is a mechanical pressure-powered type, in which an auxiliary steam supply automatically takes over to provide the motive power to discharge the condensate when stall occurs. If the steam space pressure is higher than the backpressure, condensate passes through the pump body to the float trap, which allows the condensate to discharge.

This method is more practical and economical on larger installations; for example, those using condensate drain lines of 40 mm or more.

Fig. 13.8.3 Combination pump and steam trap method of dealing with stall

2b) A pump-trap with constant flow heat exchanger (see Figure 13.8.4)The secondary flowrate does not change as it passes through the heat exchanger, consequently the steam pressure changes relative to changes in the secondary inlet temperature. At high loads the steam pressure will be higher than the backpressure, but at low loads it will be lower.

This method uses a pump-trap device, which offers the functions of a pump, steam trap and check valves in one body.

The Spirax Sarco APT14 automatic pump-trap is designed to occupy a minimum amount of space, and can be fitted to heat exchangers with nominal heating capacity of up to 1.5 MW.

It is most suited to installations with condensate drain lines up to 25 mm, but can be used on drain lines up to 40 mm in some circumstances.

A typical installation is shown on Figure 13.8.4.

Fig. 13.8.4 Pump-trap method of dealing with stall

2c) A pump-trap device with varying flow heat exchanger (see Figure 13.8.5)This method is similar to 2b), but the secondary flow through the heat exchanger varies with the heat load, due to the action of the secondary mixing valve.

The heat exchanger delivers a constant temperature water flow which is blended by the secondary mixing valve according to load. As the secondary flow varies, the steam pressure changes to maintain a constant outlet temperature, such that, at high loads, it is above the backpressure, and at low loads it is below.

Fig. 13.8.5 Pump-trap method of dealing with stall

3. Installations which ensure the steam pressure is kept constant and can never drop below the backpressure, and that the condensate can drain to and from a steam trap:

3a) Steam trap with temperature control valve in secondary circuit (see Figure 13.8.6)This method requires temperature control to be carried out with a 3-port mixing or diverting valve in the secondary circuit. The steam supply to the heat exchanger is held at a constant pressure (usually less than 1 bar g) with a pressure control valve, and as such, condensate can always be cleared from the heat exchanger against a lower backpressure.

This method is not always practical or possible. It is unsuitable on steam/air heater batteries or liquid systems where the secondary system is at such a low pressure that it is unable to prevent the liquid from boiling.

Like all methods, it has both advantages and disadvantages, which must be assessed before an option can be chosen.

Fig. 13.8.6 Constant steam pressure - Secondary temperature control

3b) Steam trap and modulating valve in condensate drain line (see Figure 13.8.7)Condensate drainage is achieved with a modulating valve in the condensate drain line. This method also maintains the desired steam pressure in the steam space regardless of load conditions.

However, it encourages (instead of eliminates) waterlogging in the heat exchanger, as control is achieved by deliberately flooding the steam space with condensate as the load reduces. Usually this method is only considered if:

The heat load is steady or changes very slowly.

The heat exchanger is designed to withstand the effects of waterlogging.

The likely stratification of temperatures of the secondary fluid is acceptable.

Fig. 13.8.7 Constant steam pressure - Condensate level control

On/off control should not be used with heat exchangersAn on/off temperature control valve does not modulate depending on heat load, but is either fully open or fully closed. An example would be a solenoid valve. When open, full steam pressure will be maintained in the heat exchanger to clear the condensate against the backpressure. At first glance, this method of control would seem to overcome any backpressure problems, but is not recommended on processes such as heat exchangers, where the secondary fluid has to be heated to its required temperature as it passes through. There are three main reasons for this:

An 'on/off' control system is activated by a thermostat which relies upon a product overtemperature to achieve control. As steam has high heat content, a significant amount of heat can be held in the steam space after the solenoid valve has shut. The overall effect is a higher product temperature than required. Should the thermostat setting be lowered to counteract this effect, the 'on' temperature may be lower than the system parameters may require. It can result in poor control of the system temperature and the potential for product spoilage.

The continual and rapid changes in pressure and temperature will impose thermal and mechanical stresses upon the heat exchanger which will probably reduce its service life.

It is never a good idea to subject steam systems to an instantaneous increase in pressure. Any condensate present in the steam space and condensate pipe is instantly pushed, by the sudden inrush of steam, through the system towards the steam trap. This can cause waterhammer, and damage the heat exchanger and steam trap.

On/off control is normally only suitable for 'non-flow' or 'batch' type heat exchange processes, notably tanks with robust heating coils, or jacketed pans, where the desired steam pressure is applied over a long heating up period (usually over many minutes or even hours). The rise in product temperature is much slower than that experienced with flow-type systems that are expected to heat the product in the short time it takes to pass through a heat exchanger.

ConclusionThe most suitable type of steam trap for heat exchange equipment in general, and especially if stall is likely, is a ball float steam trap with integral balanced pressure air vent.

If there is any likelihood of stall, a pump-trap is generally the most effective way of dealing with it, as it benefits from being:

Simple.

Cost effective.

Compact.

Please note: The diagrams in this Tutorial are schematic only, and for simplicity do not contain all the ancillary equipment that would be necessary or advisable for a specific installation. The exception is Figure 13.8.8, which shows a detailed, actual, installation of an APT14 automatic pump-trap.

Fig. 13.8.8 Detailed installation of a pump-trap with plate heat exchanger