Two-stage evaporative cooling made easy
Two-stage evaporative cooling is the most energy-efficient of all air conditioning systems and the simplest to maintain. Why doesn’t every building in South Africa have a two-stage evaporative cooling system?
By: Toon Herman
Evaporative cooling in office applications is not for the faint-hearted. Designers, developers and lending institutions are not inclined to take unnecessary risks. It took the pressure and insistence of a long-term tenant to kick-off the system in South Africa and since then a lot of experience has been accumulated.
A brief history
During the 1980s and 1990s two-stage evaporative cooling was predominantly used in industrial and retail applications and also as a pre-cooling stage for conventional air conditioning.
“But studies indicated that the system could also be suitable for the air conditioning of purpose-built, specially designed, energy-efficient buildings and some designs were prepared to test this,” Toon Herman, who designed the first office application of two-stage evaporative cooilng, explains. However, Herman continues “They were never implemented because in the end, the lending institutions considered it safer to stay with console units.”
As far as institutional office buildings are concerned, the use of two-stage evaporative cooling was never seriously considered, according to Herman. In the late 1990s however, a prospective tenant for a building in an office park stated his only requirement as 'a lot of fresh air' and the option of being able to open doors and windows. “This was the ideal opportunity to propose a two-stage evaporative cooling system,” Herman says.
“The system was quite unsophisticated. There was no provision for return air and operation was only demanded when cooling was required,” Herman explains. “Therefore the system was off during winter and heating was provided separately. Control was very coarse. A couple of space thermostats, averaged, controlled the two evaporative cooling units.”
But the inside conditions were remarkably comfortable, with temperatures similar to conventional air conditioning and a freshness and airiness that can only be achieved with a high flushing rate of 100 percent outside air. “This however had the effect of raising the expectations of the users and very soon it was forgotten that the system was not much more than a glorified ventilation plant,” Herman remembers. “Its performance was judged against that of a conventional central air conditioning system costing three times as much.”
From then onwards, two-stage evaporative cooling has slowly moved into the commercial, institutional and government building market as a viable alternative to conventional central air conditioning.
The system itself has evolved. It is now far more sophisticated, with proper supply and return air arrangement; variable volume air terminals with heating; and individual room temperature control, Herman explains. “The only difference with its conventional air conditioning cousin is that it uses no mechanical refrigeration for cooling, but gets its cooling ‘for free’ out of the ambient air.”
This also means the capital cost has gone up and is now only marginally lower than the cost of a conventional air conditioning system.
Single-stage versus two-stage
Readers should know and understand direct evaporative cooling. Consider figure 1, a wet pack with a circulating water system as an example (Pretoria altitude and typical conditions). Evaporation of water into the surrounding air cools the air. The heat required to evaporate the water is drawn from the air. This results in a drop of the dry-bulb temperature of the air. The dry-bulb temperature approaches the wet-bulb (WB) temperature of the entering air, while the air enthalpy remains unchanged. Ignore the energy inputs from fan and pump.
Now consider what happens if a cooling coil is added in front (see figure 2). Generally nothing much has changed. The total energy content of the air remains the same. The air is cooled in the coil, but exactly the same amount of energy is returned in the wet pack. But something interesting has happened.
Look at the unit-leaving conditions from top to bottom. The leaving enthalpy gradually drops from 83,4 to 49,9 kJ/kg and dry-bulb temperature from 24,9°C to 16,3°C. The top 40 percent has a higher enthalpy than the entering air and has therefore absorbed heat. The bottom 60 percent has a lower enthalpy than the entering air and has been cooled.
The top 40 percent can now be called the ‘cooling tower’ and the air, with its heat content which is called ‘process air’, can be rejected. The bottom 60 percent, at an average dry-bulb (DB) temperature of 17,3°C, has a useful cooling capacity and can be sent to the space. This is called the primary or product air.
A lower product air temperature can also be chosen at the expense of rejecting more of the original air quantity. For example, by only retaining the bottom 30 percent, a supply temperature of 16,5°C is achieved.
Note that the above is a somewhat unusual way of looking at two-stage evaporative cooling. The conventional story goes as follows: With the above ambient conditions, water is cooled in a cooling tower to 21,4°C. This water is then used in a cooling coil to cool ambient air from 32,8°C DB/19,6°C WB to 23,9°C DB/16,9°C WB. This is the first, or indirect, stage of cooling. From there, the air is then sent to a wet pack for adiabatic cooling to 17,6°C. This is the second, or direct, stage of cooling.
Why evaporative cooling?
Two-stage evaporative cooling was first introduced to save energy. Air conditioning has always been perceived as being the biggest energy user in a building, and cooling accounts for the largest part of that energy use. It must therefore be of great benefit to get that cooling part ‘for free’.
But today, air conditioning is no longer the energy guzzler it used to be. In a modern well-designed building, air conditioning only accounts for about 20-25 percent of the total energy consumption, which is more or less on par with the humble plug load – the new elephant in the room.
And conventional mechanical cooling has also become more efficient. In a typical simulation exercise, the most efficient system – water cooled chillers with spray economisers – consumed 40 percent of the air-conditioning energy, with fans and heating each 30 percent. Air-conditioning used 26 percent of the building’s energy. With two-stage evaporative cooling for the same building, the total energy consumption goes down by about 10 kWh/m2 per year (from 134 to 124) and air-conditioning now accounts for 20 percent of the total consumption. The breakdown is as follows: ‘free’ cooling, 15 percent; fans, 63 percent; and heating, 22 percent. Fans are now, by far, the biggest energy users.
Another good reason for going to evaporative cooling is the simplicity of the system, its reliability and ease of maintenance. Although most systems now have sophisticated individual room temperature control, we are still not convinced this is necessary.
Ideally, it should be seen as part of the building, unobtrusive in the background and providing basic comfort control for normal office applications. At the same time, it should allow the architect to go for a more open design in contact with the landscaped environment around the building.
But is it real air conditioning? Air conditioning is defined as a system that provides an atmosphere with controlled temperature, humidity and purity at all times, regardless of weather conditions. Based on this definition, two-stage evaporative cooling does not comply. But neither do most of the current conventional air-conditioning systems.
Maybe a more appropriate definition would be, ‘systems designed to regulate ambient conditions within buildings for comfort or industrial processes’. This definition does not specify that mechanical refrigeration must be used to create cooling, only that the system must deliver conditions suitable for the use of the building space.
Figure 3 shows a standard unit, most commonly used for comfort applications. This configuration, with return air fan; return air plenum; and mixing plenum, is used with a conventional central variable volume air conditioning system. The supply-and-return air fans are variable volume and controlled by the supply air riser static pressure.
The intention is to control the supply air temperature as close to constant as possible. The unit shown above is the ‘compact’ unit. The cooling tower pack (top) and the direct evaporative cooling pack (bottom) form one continuous unit with a water distribution section at the top and a sump with circulation pump at the bottom.
The compact two-stage evaporative cooling unit consists of the following:
Filter bank: Covers both the coil intake (primary or product air) and the cooling tower intake (secondary or process air).
Cooling coil: This is the first step of cooling (indirect evaporative cooling).
Direct cooling evaporator pack with sump and circulation pump: On top of it, and forming part of it, is the cooling tower pack with the water distribution set.
The cooling tower fan/axial (secondary air).
The supply air fan (primary air).
The return air fan plus the plenums and dampers have been introduced to adapt the two-stage evaporative cooling unit to comfort air-conditioning in a large central system where traditionally an all-air system would have been used, with chilled water as a cooling source.
It is the function of the air handling unit to deliver air at a constant temperature and a constant riser pressure. As long as the outside air DB temperature is above the supply air temperature set point, for example at 17°C, the unit runs on full outside air and the pump and cooling tower fans are switched or controlled to maintain the set point. When the outside air DB drops below the set point, the pump and cooling tower fans are off and return air is mixed with the outside air to achieve the required supply air temperature.
A constant riser pressure is maintained by controlling the speed of the supply air fan.
Note that there is no heater bank in this unit.
Psychrometrics and heat balances
Something interesting happens here:
After absorbing 117 kW of cooling over the cooling coil (work it out), the primary air actually returns 43 kW of that in the humidifier pack (second stage), for a net cooling capacity of 73 kW. (This is also exactly the amount of heat rejected by the cooling tower – as it should be.)
There is a sensible cooling from 21,6°C to 17,9°C in the second stage, but this is not an adiabatic cooling process:
The air returns 43kW of the 117kW cooling from the cooling coil. The humidifier pack actually acts as a second stage cooling tower: It keeps on cooling the water by another 3,5°C. This brings the water temperature down to 17 °C and this increases the cooling coil capacity to 117kW. But it is only borrowed: The air has to return a portion of it in the humidifier pack.
Net Room Sensible Cooling Capacity (NRSCC) with fan heat pick-up: 1°C and room design temp: 23,5°C.
Supply air temperature
The above examples indicate that a supply air temperature of around 17°C - 18°C is a realistic target under the specified ambient conditions. However, one has to look at real hourly full-year weather data. Looking, for example, at 2007 for Pretoria and Johannesburg and trying to control the supply air temperature at 17°C, there will be a certain number of hours when this temperature can not be achieved.
The number of hours, out of 3024 operating hours per year, the design supply air temperature of 17°C is exceeded:
It is clear that Johannesburg is ideally suited for two-stage evaporative cooling, more so than Pretoria. Most of the interest comes however from Pretoria and from the above figures it would make sense to design a Pretoria building for 18°C supply air and Johannesburg for 17°C.
However, this would mean that Pretoria would require 20 percent more air for the same load. This might not be necessary as the following simulation results will show.
For some of the time, the supply temperature will be above 17°C and this will affect the room temperature.
Investigating this is part of the simulation process.
Room temperatures achieved
It is the intention that a two-stage evaporative system achieves the same room temperatures as conventional air-conditioning. The following results are for a typical four-storey building, of average thermal efficiency, with a system designed for an internal temperature of 23°C with supply air leaving the unit at 17°C for both Pretoria and Johannesburg. The room temperature is controlled at 22°C ± 1°C.
Table 2 is an extract of the summary sheet of the temperature statistics for some of the rooms.
As can be seen, a supply temperature above 17°C does not necessarily have a proportional effect on the room temperatures.
The worst room, CB_SO exceeds the design temperature of 23°C for only 7,4 percent of the time in Pretoria, 2,3 percent in Johannesburg and stands out from the rest of the building. It is clear that the façade or the system serving that room should to be looked at again. This simulation process is a dynamic tool for the design team, allowing for adjustments to increase overall performance.
The term ‘discomfort index’ requires some explanation. It is an area-weighted indicator of how much and for how long the design temperature of 23°C is exceeded by on average throughout the building.
For example, a discomfort index of 25 means, on average, the room temperature exceeds 23°C by 1°C for 25 hours per year, or by 0,5°C for 50 hours per year, and so forth.
Relative humidity is of concern with evaporative cooling and should be checked as part of the simulation process.
For the same building as above, the statistics for the average relative humidity are as follows:
According to ASHRAE, it is generally believed that the relative humidity in office space should not be higher than 60 percent. However, we have experienced relative humidities of 75 percent and higher in an office environment and that was quite acceptable. (The temperature was 22 -23°C).
From ‘An investigation of thermal comfort at high humidities’ by Mark E Fountain et al the following extract:
“McIntyre (1980) cites several studies showing that for operative temperatures within the comfort zone, differences in RH as disparate as 20 percent and 70 percent can be undetectable, let alone a source of discomfort.”
Another potential problem with high humidity is mould growth. It is difficult to find relevant information on this subject. Most studies refer to winter conditions in the northern hemisphere and the growth of mould on surfaces on or near thermal bridges in the outside walls. A humidity level of 80 percent is mentioned but it is not clear if this applies to our conditions.
Energy and water consumption
See Table 4 for a comparison with conventional air conditioning. “The most energy efficient conventional system uses a water-cooled chilled water plant with a spray-based economiser cycle,” says Herman. A more common conventional air conditioning system employs air-cooled chillers and has no economiser.
The systems simulated above are of the ‘orthodox’ variable volume type, namely a constant supply air temperature, throughout the year, and for the total building, regardless of zone and orientation. This was the traditional configuration, when variable volume was first used, 40-50 years ago. A variable volume system however presents the designer with a dilemma: How to deal with the minimum air required for ventilation and for the delivery of the heating under low load conditions? In the early days, there was always a generous lighting load (up to 40 W/m2) and even later, when the lighting dropped to 20 W/m2, there was still sufficient load to deal with the minimum air at constant supply temperature.
But lighting is now far more energy-efficient (as low as 5 W/m2) and can no longer be counted on to disguise the minimum air reheat problem, especially when day-light and occupation sensors are employed to reduce the lighting load even further. As can be seen in Table 4 heating energy consumption is substantial and a big chunk of that goes to minimum air reheat. Two-stage evaporative cooling does a lot better than conventional cooling because the supply air temperature is 4 - 5°C higher and therefore requires less reheat.
Designers have tried to reduce this waste of energy by zoning the building and adjusting the supply air temperature to suit the requirements of the zone. This of course assumes all rooms in a zone have similar load profiles. It also involves the use of sophisticated control systems to somehow pick out the room in a particular zone with the highest cooling requirements and adjust the supply air temperature to suit that room.
To fairly compare the energy performance of the two systems, we will ignore the heating consumption as this should be the same for all once the minimum air reheat is resolved.
As can be seen, ‘free cooling’ comes at a price. In the above example, the price is a higher air quantity and therefore higher fan energy consumption. But the difference is still 25 percent compared to the best conventional air-conditioning system. Surprisingly the water consumption is also lower. At the other end of the spectrum, the air-cooled chiller system with no economiser, which is quite common, uses more than double the amount of energy.
The differences (6 to 30 kWh/m2 p.a.) may appear to be small when looking at total building consumptions of 200 to 400 kWh/m2, but they become significant when trying to achieve ultra-efficiency targets of 100 to 115 kWh/m2 per year for the total building consumption.
In the case of the above simulation (Pretoria), the system runs on full outside air for 77 percent of the time. Return air is introduced when the outside air DB drops below 17°C but the outside air supplied never drops below 11 l/s per person.
South African application
Two-stage evaporative cooling originated overseas, but South Africans took it one step further. “We know of no examples overseas where the system is used in high grade office space as we do here,” says Herman.
Two-stage evaporative cooling started off in South Africa as a glorified ventilation system, with very coarse control. It was simple and inexpensive. However, it very quickly evolved into a system that could compete with conventional air conditioning, with the same type of controls, and with the same expectations from the occupants. One such project, and the largest to date, is the Department of Trade and Industry (DTI) campus in Pretoria.
Energy savings for commercial buildings in Europe
Case study: Dynamic transient thermal simulation, undertaken by Lennox, for the mall of the Carrefour Pontbault Combault commercial center close to Paris.
Presented by Steve Kimber and Katharine Haarhoff
Faced with challenges relating to competitiveness and climate change, Europe has launched several initiatives for an intelligent energy strategy. New regulations reflect the European Union’s commitments on climate change and its determination to reach 20 percent efficiency improvement in buildings by 2020.
One of the major requirements of the new initiatives is to issue energy certificates for buildings: Energy Certificates grade the energy efficiency of a building based on its energy consumption. It is believed that soon energy certificates will affect the value of a building.
What is building thermal balance:
A building is a ‘closed’ system where the amount of energy required to maintain a set temperature will depend on internal and external thermal inputs. (See figure 1)
The gains and losses are put into a mathematical model called an L3C simulator. L3C is a transient dynamic thermal simulation tool designed to compare and value potential energy savings on commercial building through unit or building improvements.
1: There are two ways to save energy in commercial buildings.Improve energy efficiency of the building itself by:
2: Improve ‘dynamic’ energy efficiency through better air conditioning solutions:
Improve insulation, reduce air leaks, and thermal bridges;
Improve windows, sky domes, and solar protection; and
Change temperature set points between summer and winter.
Use the free cooling;
Use a new generation class ‘A’ rooftop;
Use variable speed drive on the supply fan;
Improve fresh air management; and
Use heat recovery on exhaust air or food refrigeration systems.
The Lennox team looked at this simulation with a view to maximising energy savings. They looked at the impact of:
Pressure drops increase;
Setting unoccupied periods and stopping the fan during dead zones;
Changing temperature set points;
Installing a class A new generation Baltic ™ Rooftop unit;
Usine eDrive™ variable speed with direct drive supply fan; and
Managing fresh air input.
Initial Building Characteristics
(The starting point)
What is the current position?
This first calculation illustrates the annual energy consumption of an old building. The building has low insulation and a high leakage rate; is equipped with old R22 rooftops, with low efficiency fixed airflow, which is not set properly – a no fan stop in DZ; and no specific unoccupied mode settings.
Building thermal loads:
Estimated annual energy consumption for HVAC: 577,700 kWh/90,6 kWh/m²/year ® around R408 000 (€40 440) per annum.
95 W/m² in summer
67 W/m² in winter
HVAC energy consumption split:
Air treatment represents the biggest potential for energy savings with 57 percent of the total annual energy consumption in this case.
To save energy on ventilation we can:
Reduce operating hours for the fan by:
Setting unoccupied periods; and
Stopping the fans during dead zones.
Reduce airflow rate by:
Adjusting airflow rate depending on the thermal load and fresh air needs.
Limit pressure drops by:
Regularly carrying out maintenance of filters and diffusion; and
Limiting usage of heat recovery modules as they can increase pressure drops a lot.
Impact of more pressure drops
Poor maintenance of filters, ducts, and diffusers; and
Poor sizing of the ductworks.
Impact on annual energy consumption:
An increase of 150Pa on pressure drop on the system of this case study can cause an increase on annual energy consumption from 13 to 15 percent depending on the age of the installation.
On existing equipment badly set without a dead zone or unoccupied mode:
+ 73,491 kWh/year + approximately R51 000 (€5 144) per annum (+13 percent)
On new equipment correctly set:
+ 25,418 kWh/year + approximately R18 000 (€1 779) per annum (+15 percent)
The energy consumption due to the supply fan is then increased to 63 percent of the total annual energy consumption for HVAC on this building.
Impact of setting unoccupied periods
On this installation, unoccupied period represents 4076 hours (47 percent)
Definition: Unoccupied periods correspond to the hours where the building is closed and has no need for fresh air or to maintain comfort temperatures for the occupants.
Impact of the wrong setting of unoccupied period on annual energy consumption:
On existing equipment:
- 77 440 kWh or – approximately R54 000 (€5 420) per year (-13 percent)
On new equipment:
- 16,800 kWh or – approximately R12 000 (€1 200) per year (-4,3 percent)
Impact of stopping ventilation during dead zones
Here, dead zones in unoccupied period = 3 495 hours (40 percent)
Definition: The Dead zone corresponds to the periods where the ambient temperature is between the winter and summer set points and there is no need for cooling or heating.
Impact of stopping the fan during dead-zone in the unoccupied periods:
Cumulated gains of setting the unoccupied period correctly and stopping the fan in dead zones:
On existing equipment:
- 222 473 kWh or – approximately R157 000 (€15 573) (-39 percent)
On new equipment:
- 144 600 kWh or – approximately R102 000 (€10 120) (-37 percent)
Impact of changing the set points:
Temperature set points in commercial buildings are usually set to 20°C in winter, and 24°C in summer – or less during occupied periods such as with the initial calculation of this study.
- 52,300 kWh / - approximately R36 000 (€3 660) (-15 percent)
‘Legal’ temperature for winter operation: 19°C
‘Legal’ temperature for cooling operation: 26°C
Potential energy gains from changing set points from 24°C to 26°C in summer and 20°C to 19°C in winter:
Cooling: -36 percent (-38 980 kWh per annum)
Heating: -21 percent (-13 750 kWh per annum)
Cumulated gains after correct setting of the existing units
Impact of replacing HVAC equipment
Packaged air conditioning systems operate only four percent of the time at full load according to prEN14825.
New rooftops such as Lennox Baltic™ are Eurovent class A (Net EER › 3.0, Net COP › 3.4) and feature new refrigeration circuits designed to provide best part load efficiencies.
Multi-circuit tandem compressor assemblies
Increase heat exchange area
Alternate defrost thanks to dual circuit from size 045
288,852 kWh = - approximately R25 000 (€2 527) (- 12 percent versus existing rooftops)
Cooling: - 18 percent (- 12 225 kWh per year)
Heating: - 5 percent (- 2 757 kWh per year)
Ventilation: - 12 percent (- 2 125 kWh per year)
The reduction of energy consumption of the fan for the same airflow is due to the increase in the box size = less pressure drop.
eDriveTM variable speed drive
This changes the mix of cooling, heating and ventilation (see before/after)
The fan motor still represents more than 60 percent of the annual energy consumption for our building.
eDriveTM can help reduce energy consumption of the fan thanks to variable airflow rate control.
eDrive™ variable speed drive
Airflow rate reduction during part load operation and dead zone;
Friction-loss elimination thanks to direct drive transmission;
Accurate airflow rate setting during start with eFlow™;
In this case study the airflow rate is reduced between 100 percent and 75 percent of the nominal airflow rate during part load and dead zones.
Impact on annual energy consumption:
- approximately R51 000 (€5 083) per year (-27 percent versus previous calculation)
Cooling: - four percent (-1 974 kWh per annum)
Heating: +18 percent + 8 660 kWh/an
Ventilation: -49 percent 79 300 kWh/an
The average power input to the fan is reduced:
NET cooling capacity increases ? The energy consumption in cooling is reduced in the same way the heating consumption is increased as it must compensated for the reduction in heat input from the fan.
Intelligent fresh air control depending on occupation
Here we propose to adjust the fresh air rate to the actual occupancy level, through CO2 sensing:
Introducing fresh air in a building is mandatory.
New certified buildings are “air tight: and require better management of fresh air input to maintain air quality.
Fresh air is usually set on max design occupation varies along the day and the week.
Total annual fresh air volume can be reduced by up to 72 percent (Hypermarket application).
Approximately -R16 000 (-€1 643) per year (-12 percent versus previous calculation)
Impact of the modifications on the building load
Max cooling load: 601 kW ® 478 kW
Max heating load: 421 kW ® 283 kW
Savings totalling 407 000 kW/hr have been achieved in six steps:
Scheduling the cooling/heating off during unoccupied periods;
Stopping the fan during the dead zone in unoccupied periods;
Changing the set points (19°C heating and 26°C cooling);
Replacing the equipment with modern efficient units;
Controlling the supply fan using variable speed; and
Regulating the fresh air using carbon dioxide monitoring.
Humidification and de-humidification, part one
The above terms simply mean the addition and removal of moisture or water vapour from the air. For two major reasons, environmental and enclosure conditions, these processes are important and we look at reasons of effecting them in the air conditioning installations. Eventually the effects can lead us to be able to measure relative humidity of air for specific occupancy.
1. Environmental conditions
2. Enclosure characteristics.
Human comfort:During cooling in air conditioning applications, we know air reaches a dew point temperature. That means vapour is removed from the air and condensation is realised. If the air temperature becomes well below human comfort, it increases evaporation from the nose membrane and throat and causes skin and hair to dry out. So the relative humidity in areas for human comfort should be above 20 percent and less than 60 percent.
Process control and material storage:The following factors are controlled in the space for processing or storage of certain products: regain; rate of chemical reaction; rate of bio-chemical reaction; rate of crystallization; product accuracy and uniformity; corrosion; rusting and abrasion; static electricity; cleanliness; and product formability.
Static electricity:Electrostatic charges are generated when materials of high electrical resistance move against each other. The accumulated charges may result in sparks; difficulty in handling sheets of paper; clinging of dust to oppositely charged objects; and explosive gases. If moisture content of the air is increased the charges can be controlled.
Prevention and treatment of diseases:At 50 percent relative humidity, the mortality rate of certain organisms is highest and the influenza virus loses much of its virulence. The mortality rate decreases both above and below 50 percent relative humidity. Operating theatres in heath facilities are critical.
- Sound absorption is higher in lower humidity, above 15 percent and below 20 percent. Remember humidity control becomes effective if the temperature is well controlled.
Normally in certain countries we experience low humidities during winter. The extent of adding moisture to the building will depend on its walls, roof and other elements to prevent or tolerate condensation. The visible condensation may result in deterioration of the surface finishes, mould growth and a reduction of visibility through windows.
During the building and air conditioning design process, the relative humidity should be taken into consideration.
To be continued in the next issue with part two.
 RACA, ASHRAE
Additional thoughts on cooling control, part 4
In this series we have been mentally exercising on control issues, and more specifically on the control of a chilled water application with cooling, on one hand, by a reciprocating compressor having four steps of capacity control. On the other hand, we looked at achieving this amount of cooling, and setting up control accordingly, with four scroll compressors being applied this time around. Early in this series we looked at control achieved from a temperature sensor placed in the water approaching the chiller, and then later at the circumstances where the temperature sensor was to be placed in the leaving water.
Our method was to base the thoughts on an electromechanical step controller, for the simple reason that we could picture the motor-driven cams being turned, and the micro-switches, with their cam-following fingers and little wheels riding up and down to initiate their switching actions. This way, the opening or closing of the switches on a pre-determined basis in response to the controller’s commands was how we achieved the programmed control. The cardinal issue so far has been that control of such a set-up may be based on either the measurement of entering water temperature, or upon leaving water temperature. Both have their virtues and both have their downsides, as we carefully considered.
The critical warning was that the chosen arrangement had to be specifically configured. If we hauled the sensor from, say, its position in the entering water and instead placed it in the leaving water because that made ‘more sense’ and attempted some minor programming adjustments of the controller to ostensibly accommodate this move, this would produce chaotic operation, frequently leading to failed compressors. Unfortunately, not everyone has been able to recognise chaotic operation when they encounter it. Typically, a large compressor would start, load up, proceed to unload and then stop, all in the space of a few minutes. Witnesses would remain unmoved. I am able to relate several personal experiences where exactly this has happened. The big alarm bell was that the people involved in most cases never became aware of the nature of the major error that had been in-built. They would replace or have repaired the failed compressor, and often carry on exactly as before, only to be greatly dismayed by yet another compressor failure after a surprisingly short time.
I earnestly hope all our readers have grasped the subtlety of this perilous situation, and will never get caught out in the way of so many of our unfortunate predecessors. While, to this point it has been alright to lump four individual scroll compressors into the same ‘block’ as those of a single reciprocating compressor with multiple capacity stages, the fact is that it’s not really acceptable. There is indeed a further difference.
The issue is, the reciprocating compressor itself wouldn’t be stopping and starting for every change in required capacity. The machine would have some form of capacity controller, sometimes hydraulically actuated and most commonly responsive to operation of a solenoid valve. This function would determine whether an operating piston and cylinder combination should be rendered either active or idle. Usually the targeted suction valves are physically held open while in the ‘unloaded’ situation, thus disallowing compression in the cylinder or cylinder pair for this period. Such an arrangement may load and unload repeatedly at very short intervals without experiencing any harm.
Not so, a standard scroll compressor – which stops or starts to respond to a requirement to match a change in the load of the moment with a change in capacity requirement – must provide a warning signal with each step change brought about by a motor either starting or stopping. It is with this thought in mind that we move on from where we left off last month. We will think of common failures in scroll compressors.
What lurks behind failures?
In my various exchanges on failures of scroll compressor difficulties, it has become increasingly apparent that a compact sequence of short runs offers a huge threat to the well-being of such a compressor. The rationale that underlies this is well worth further emphasis. Across those first seconds of operation which follow any start, such a compressor will throw an amount of oil. The anticipation is that, as the machine continues to operate over the next several minutes, this displaced oil will creep through the evaporator and then progressively migrate back to the compressor by way of the suction line. In this way, a normalised situation is quickly restored in the oil charge.
But, if there is no worthwhile ensuing run following the start, and the compressor is very soon again stopped, that displaced oil will be temporarily lost from the compressor. One or two such cycles possibly will not displace enough oil to cause the compressor any grief. The trouble will come if circumstances militate to result in this potentially harmful sequence consistently repeating across an extended period of time.
A useful solution could be to have one compressor of the group capacity controlled. Then the controller must have the capability of always repositioning operation of this compressor within the overall pecking order, so that it can provide the necessary in fills to offer infinite capacity control at whatever point it may be needed across the working capacity range of the entire group. In other words, while system load is in the range of zero to 25 percent, this compressor alone will run. But, if load rises to above 25 percent, a fixed capacity compressor will be started, and the variable capacity machine will back off to provide less than five percent of the installed capacity. This will be repeated as load climbs or reduces through all the programme steps. In this way, a relatively smooth ramp of capacity is offered throughout almost the entire range. It just requires much skill in applying the correct control hardware and adjustments. Those details, as best as I can configure them, will come next month. We first have other issues for discussion!
Figure 1 illustrates diagrammatically how we would hope to see this objective. To the best of my knowledge, there are just two ways by which this capacity variation function may be secured. The most obvious is to apply an inverter variable frequency control arrangement.
Application of an inverter makes it possible for a squirrel cage motor, which normally operates at a constant rotational speed, to be driven across a range of speeds. It does this by changing the frequency of the current being fed to the motor, on a controlled basis. But in most cases a compressor designed for fixed rotational speed is not suited to being subjected to this treatment. When a compressor has been manufactured specifically for such an application, it will probably have been designed to operate across the frequency range of 80 Hz down to 30 Hz, for example. It will be supplied with a suitable inverter as ‘part and parcel’ of the package.
But a difficulty can arise from the fact that the inverter is not likely to be as pristine as a Rolls Royce. Given ‘dirty current’, a ‘run-of-the-mill’ inverter is subject to a risk of failure. Back when I still had a day job, one of my periodic tasks was to record voltage of the electrical feed coming into some plant we might have had under review. In cases where a plant was suspect, I occasionally had to run across the roof of a building while dodging lightning bolts in order to set up the voltage recorder so we could later see how the installation was being affected by the storm.
Thunder storms in our parts, back when I was actively doing these things, seemed to occur considerably more frequently and with substantially more spectacle than is presently is the local case. When I was at school, we counted on getting two hefty thunder storms for every three days throughout the month of February. While I still live near enough to the same spot, an entire February can now come and go without presenting us with a single thunder storm. Under stormy conditions, it was common to obtain a recording very much as has been simulated in Figure 2. This would be garnished with voltage spikes, indicating how the electrical feed had been affected when there had been lightning strikes nearby. And those recorded spikes were plentiful across the hour or so duration of the storm. It would play havoc with some of our plants, although with the rugged and unsophisticated equipment of the times, seldom causing anything worse than a ‘nuisance trip’.
Copeland’s ‘digital scroll’
From what I am told, it is ‘ordinary’ inverters that can succumb to a feed which has become ‘spikey’, and can occasionally go belly-up for this reason. This susceptibility is the reason Copeland state it offers its ‘digital scroll’. They claim a machine using Copeland technology costs less than does an inverter-controlled machine, while offering greater capacity range and durability than the inverter option. It will be productive for us to take a look at Copeland’s technology in this regard.
Figure 3 illustrates the concept of the digital scroll. In the standard scroll compressor, the scroll tips of the orbiting and fixed scrolls are pressed into constant snug contact against the inner surfaces of the opposite scroll. Thus the machine provides fixed capacity when capacity is viewed in terms of the variables of suction and delivery pressure. Apart from going to a variable speed machine, there is a possibility that only a digital scroll to offer the means to alter capacity to match changes in refrigeration load.
The modified components applied to achieve this include first and foremost a fixed impeller which has been provided with the ability to be minutely floated up and down with a travel of a millimetre or so; just enough to free up contact between the scroll tips of the fixed and orbiting impellers. If we picture the entire situation here, there is concentric compression taking place as gas is ushered inward by compressor operation along its spiral path of diminishing volume from its two opposite points of entry at the periphery, to its ultimate compressed state at the central delivery port. For this to be achieved, snug contact is required between the scroll tips and their opposite contact points throughout orbital operation. However if this contact is broken, even when a clearance as small as just a millimetre occurs, pressure will equalise right across the impeller. As there is a check valve, or a one-way valve or non-return valve if you prefer, located in the delivery port, high pressure gas that has already been delivered is prevented from rushing back into the impeller.
How this is achieved
The fixed impeller has been equipped with an arrangement which contains some of the genes of a hydraulic relay. At the head of affairs is a form of piston. Compressed vapour is fed by way of a restrictive drilling, and is free to distribute across the face of the piston. Figure 3 makes a valiant attempt to show this.
There is also a bleed arrangement from the cylinder of this sub-arrangement. This bleed passes through a solenoid valve, and then returns to the suction line. Two things can happen:
If the solenoid valve is closed, there is no draining away of the feed of gas that has passed through the small drilling from the high pressure side of the scroll. This pressurised gas spreads to fill the piston area, causing the arrangement to be thrust down. This closes the contact of the scroll tips, and the compressor functions normally. This closure is done by way of a ring of compression springs, so as to get the contact pressure ‘just right’.
However, if the solenoid valve is opened, vapour in this leg will freely vent to the low pressure environment of the suction line. As vapour is being freely bled from the downstream side of the restrictor drilling, the entire region will fall to a pressure almost as low as suction pressure. With the previous down-force now having been removed, the impeller will be free to rise, pressed upward by gas pressure from below. The small clearance we have discussed is now realised at the scroll tips. Gas which is in transit through the impeller promptly bleeds across the scroll tips from high to low. The check valve closes, and low pressure prevails throughout the impeller. Notwithstanding continued operation of the impeller, there now will be no compression or delivery of vapour to the high side. The practically free-wheeling motor will consume minimal power.
Thus, by having a control system that determines whether the solenoid valve is to be activated or de-activated governs when the machine will pump vapour and when it does not.
Figure 4 will move us forward. The ‘smart’ controller divides real time into ‘slots’ for purpose of the control. The duration of these ‘time slots’ will be automatically optimised at between 15 seconds and 30 seconds. The solenoid valve may remain closed throughout. It may be held open for the entire period, or it may switch between ‘open’ and ‘closed’ at any point within each time slot. This duration of each ‘open’ and then the ensuing ‘closed’ phase is tailored by the controller as its response to the instantaneous load. The impeller may be commanded to deliver refrigerant for as little as 10 percent of the running time, and from there, anywhere up to 100 percent. The machine cannot be called upon to work at beneath 10 percent of its capacity, for its motor will be continuing to generate heat, and it is essential that this be transferred onward in a feed of refrigerant to travel to the condenser for due release from the system. The delay of these few seconds of ‘non-pumping time’ is however inconsequential.
‘Entering’ versus ‘leaving’ water control
In this series I have done my utmost to make plain the entirely different approach that is essential, according to whether ‘entering’ or ‘leaving’ water control is to be applied. It will probably have been clear from the discussions that it was the electromechanical controls of the past number of years that have been involved in occasional pandemonium, usually due to the sensing bulb of an entering control arrangement having been misguidedly placed in the leaving water, in the fictitious belief that this would ‘make the job work better’. The reverse would be equally bad, but it doesn’t seem as attractive to finger-happy folk.
But, to bridge into the next part of today’s story, we will consider an unfolding event that took place in a town less than a million miles from here. During an out-of-town training stint, I was asked to look in at a troublesome installation. It was a major installation, serving a water-based thermal storage arrangement. This installation incorporated four chillers, each equipped with its own pair of hefty reciprocating compressors.
The compressors serving these chillers were periodically suffering fatal trauma. As part of my free service for friends, I went to where dead compressors are mourned, and inspected the stator illustrated in Figure 5. The massive burn in its windings is apparent. What happens in many a basement plant is that the plant room is placed directly alongside the building’s transformer installation. The feed cable is therefore very short, making it possible for enormous amperage to flow in the event of a severe short circuit. Had this same machine been located in a rooftop plant room and undergone a corresponding measure of short circuit, the lengthy rising cable feeding the upstairs plant room would helpfully restrict the overfeed of current. The circuit breaker protecting the motor would trip, and upon subsequent repair, a somewhat smaller amount of damage would have been revealed.
Why should the rotor windings in this case have suffered such a traumatic fault? Say maximum running current rating had been 400 amps. The circuit breaker would have been set to trip at 400 amps after a short delay at which cut-out amps would have been greater to accommodate the current inrush associated with starting. Primary protection is the affair of the motor starter, and we are presently talking of the back-up current protection. In the event of a burn, this back-up protection would come to the rescue, so as to limit the extent of damage. This ‘back-up protection’ might comprise a suitably rated three-phase circuit breaker, or it could comprise fuses.
In this case, the installation had a circuit breaker for each circuit. But this circuit breaker would only be able to protect against a certain maximum current inrush. Say it was rated to protect against a current flow of up to 2 000 amps. If we had the upstairs plant room scenario with its lengthy cable feed, this could perhaps have limited the current rush of this particular case to 1 500 amps. The 400 amp breaker would have successfully tripped, protecting the associated motor switchgear and limiting the event to a minor burn.
When in a basement
Given a corresponding setup, but in the scenario of a basement plant room with its switchboard positioned close to the transformer of the main incomer, the same fault may well have yielded a current flow of 3 000 amps. The breaker would have tripped, but would have drawn arcs spanning the air gaps, allowing this 3 000 amp fault current to continue to flow. Dependence would shift to a protective device further up stream, and therefore with a considerably higher amperage rating. Now, instead of being rapidly quenched, the fault could persist even for minutes and longer, providing adequate opportunity for producing the advanced damage we see in Figure 5.
A High Rupturing Capacity (HRC) fuse has been inset in Figure 5. When applied in the circuitry, such fuses make it impossible for an arc to draw in the event of severe overload. The outer shell is of a hollow ceramic material. This contains the copper strip which forms the fuse element passing through the centre. This is completely immersed in suitable fine sand. Should the fuse rupture, the sand rushes in to occupy the void left by the vaporised copper. This instantly quenches the arc that would have formed, immediately halting the electrical feed. The damage that would have occurred in the Figure 5 stator would have been far less, although the motor will still have required a full rewind. In the absence of such protection, contactors would have welded closed, and you could have bet your boots the switchgear serving the Figure 5 motor would have been reduced to molten scrapheap copper by the event. HRC fuses would have probably protected the bulk of the electrical gear from such catastrophic failure.
A present dilemma
Although it’s informative, it is not in itself of specific importance with regard to our present dilemma. As mentioned, this was a chilled water thermal storage installation. And as we have said, feedback is absent in a case where the water being processed through the chiller might last have seen action some hours previously. The object of the exercise is to fill the thermal storage tank with water at 5ºC. Therefore the consulting engineer had rightly specified leaving water temperature control.
Not only that, with those earlier advances of technology some years back, all services were required to ‘keep up with the Joneses’ and have everything that could be controlled be controlled from a magic state-of-the-art computerised Building Management System (BMS), stashed into a PC. Therefore the original control arrangements that were part and parcel of the original chillers were orphaned by some deft snipping with a pair of side cutters. The cyber brain would now be invested in part with that control authority of the complete chiller installation. In compliance with the specification, that would be in terms of leaving water control. The standard electromechanical control, with which each machine has been imported, was indeed conformed for entering water control. But these controls were destined to serve out their lives as never-used museum pieces in the panels of the individual chillers.
I looked at the stator of Figure 5, and also at the near-by pile of scrap metal that constituted the earthly remains of the compressor that had most recently died. I had been standing in the wrong queue when ‘computer geek’ qualifications were being issued, and therefore know next to nothing of software and its performance. However, the detritus carried emblazoned all over it the all-too-familiar message which declared ‘Entering water control has been applied with a leaving water sensor’. This had to be it. Presumably the software writer had dug up an algorithm for ‘chiller control’ from some or other source. As a non-expert in air conditioning matters, he or she was unaware of the thrust of those critical issues we are presently discussing. The entering control routine had its ‘temperature’ function therefore orchestrated from a leaving water temperature transmitter. The settings parameters would have been subjected to ‘creative’ treatment to provide some hint of respectability.
This reportedly had by no means been the first such failure of compressors within this plant. But it would appear that, till now, the diagnosis had been “Well, having compressors fail is life for you”, and the butcher’s picnic was to continue unabated. While no follow-up report was ever favoured to me, I imagine the software specialist sought and found a P + I + D algorithm that incorporated a dead zone. Someone worked out the smallest workable value for the dead zone from compressor capacity control data, and the plant has run from that day to this with no further hiccup.
The reason for these considerations
What has been our reason for having travelled the latter part of today’s road? This has been to illustrate my belief that a competent computer programmer is not by any means also a polished refrigeration expert. And I believe that to have been a great problem of that era of a decade or so back. There are excellent new control types presently available that we could not even have dreamed about back during that period. We shall speak of a few of these. Their software is truly phenomenal, thoroughly tested and is not in any sense of the word ‘home made’. It is excellently documented, and brings the achievement of outstanding control success well within reach.
But that will take us across a fair amount of additional territory, and will aid us in extending the line of thought initiated by our Figure 1. Preparing this next article took me on a visit to my old friend from many years back: the one and only Pravin Kumar. He provided me with a considerable measure of guidance. But we already have had a full house for today. So we shall reserve those discussions concerning what Pravin passed on to me for next time. All being well, this will be in October. I look forward to us again being together then.