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Optimization of Refrigeration Plant Operation. Engineering Approach

Sergei Khoudiachov

S K Energy Consulting Ltd.

Canada


Abstract

For years, many professional engineers have been focused on the design of energy efficient refrigeration plants. Although a well designed refrigeration plant is the foundation of energy savings, the efficient operation of a refrigeration plant is the ultimate goal of the energy saving process in industrial refrigeration. Efficient operation enables all of the energy savings originally designed into the refrigeration system. Ideally a refrigeration plant control should have:
• Software to determine optimum set points and optimum operating strategies
• A control system to implement these set points and operating strategies

There are several PLCs and computer based control systems on the market which can properly implement certain set points and operating strategies if they are properly and painstakingly programmed. But notably, software that is capable of automatically determine optimum set points and optimum operating strategies does not exist, as it is very complicated to create this type of software. Who then should determine these set points and operating strategies? Typically this job was given to the operators. They do their best to optimize the operation of the refrigeration plant, but often make mistakes during the process. To maximize efficiency of the refrigeration plants, optimization should be focused on major energy saving measures such as optimum condensing pressure, optimum suction pressure, optimum defrosting and etc. During periods of cold weather, condensing pressure should be kept as low as possible because optimum condensing pressure at this time can be 60 psig [4.1 bars] or lower. A major misconception in industrial refrigeration is that refrigeration plants cannot operate at condensing pressure lower than 100 psig [6.8 bars]. To lower condensing pressure below 100 psig [6.8 bars], several barriers need to be overcome. However, every barrier has at least one solution. This paper will describe an engineering approach to optimizing the operation of refrigeration plants and will analyze potential missteps in this process.

Introduction

In a modern economy, companies realize that sustainability will be a critical part of their strategic plan. Two key factors reignited the quest for sustainability; energy and climate change. Energy efficiency can have a significant impact on both of these factors and it is now a key component of wise business practices. Energy conservation is the right step for the environment and it is the right step for the business.

For years, many professionals have tried to save energy in industrial refrigeration. The design of the refrigeration plant is the major focus of this effort. Although a well designed plant is the foundation for energy savings, the efficient operation of a refrigeration plant is the ultimate factor in the energy saving process. My experience has shown that optimization of refrigeration plant operation can affect as much as 70-80% of the total energy savings. Unlike in some other industries, energy use of one component of the refrigeration plant has an influence on energy use of another component. For example, operation of an additional condenser fan will reduce condensing pressure and thus compressor energy will be reduced. Interaction of compressor and condenser energy use is a complicated function. Only a complete system approach will give us an opportunity to maximize energy efficiency of the refrigeration plant.

The control of an ideal refrigeration plant should have:
• Software to determine optimum set points and optimum operating strategies.
• A control scheme to implement these set points and operating strategies.
There are several PLCs on the market, which can implement certain set points and operating strategies. However, software that is capable of automatically determines optimum set points and optimum operating strategies does not exist, as it is very complicated to create this type of software. Therefore, the question becomes: Who should determine these set points and operating strategies? Typically this job was given to the operators.

Operators

Several years ago, a few operators of process facilities tried to optimize operation of their refrigeration plants. During these optimizations, the following steps were taken;
• Establish a system performance baseline
• Compare system performance with other plants or industry benchmarks
• Identify potential energy efficiency improvements
• Implement the best energy saving opportunities

This approach is time consuming and required hundreds of different tests. Typically it would take 1 to 2 years to do these types of tests. Many companies do not have the luxury to spend many hours of labor to test and to analyze operation of their plants. The operators mentioned above have achieved only partial optimization and some results of their optimizations were questionable. Optimization of a refrigeration plant operation can be a very complicated issue and it should be done by the experts in order to save time and money. Operators will be major players in this process, but they should have a different role.

Experts

It is thought that in the future we will have software to automatically determine optimum set points and optimum operating strategies. Thousands of tests will lead to the creation of several universal equations, which will describe operation of refrigeration plants. Specific information about a particular plant will be entered in this software and optimum set points will be determined.

Experts have a similar approach. I know several energy efficiency consultants that have done thousands of tests on different refrigeration plants and based on the information received they can determine the optimum set points and optimum operating strategies for other refrigeration plants. Using this approach, many companies would save hundreds or thousands of hours of labor and the results of these optimizations would be very close to the optimum. Operators would have the role of the testers, following the advice of the experts to tune up the operation of their refrigeration plants. Optimization of refrigeration plants does not necessarily require a large investment, but it does require the highest level of knowledge and experience. Operators can claim that they have experience in operating refrigeration plants, however, there is a significant difference between regular operation and energy efficient operation. These operations can be compared with driving a car. Regular operation can be compared to every day driving to work and back home. Every day the same route was taken and car speed was relatively steady. Energy efficient operation, on the other hand, is like car racing. To win the race, maximum effort is required from the car (plant), as well as from the driver (operator). There are a number of different ways to save energy in industrial refrigeration. However, priority should be given to the major energy saving measures such as: optimum condensing pressure, optimum suction pressure, optimum defrosting, optimized compressor sequence, use of VFDs and remote system tuning.


Key Energy Saving Measure #1: Optimum condensing pressure

Summer Operation

Industrial refrigeration plants have evaporative condensers whose capacity varies with wet bulb conditions. Today, many companies claim they “float” condensing pressure during summer operation. To optimize operation of the refrigeration plant, condensing pressure should not just to be “floated”, but it should be “floated” at optimum level to minimize energy use of the whole refrigeration plant. What is the optimum level? Modern refrigeration plants have PLCs with a wet bulb approach feature to control condensing pressure. This feature can be helpful to balance capacities of compressors and condensers to minimize total energy use. Typically, when the compressor capacity is reduced, to keep the right balance, the condenser capacity should be reduced as well.

Example
Assume that optimum wet bulb approach is 10 degF [5.6 degC]. Wet bulb temperature is 75 degF [23.9 degC]. Optimum condensing pressure will be 75 +10 = 85 degF [29.4 degC] and corresponding optimum condensing pressure will be 152 psig [10.3 bars].

To determine optimum condensing pressure, wet bulb temperature should be measured and optimum wet bulb approach should be determined. My experience has shown that at wet bulb temperature of 70 degF [21.1 degF] or higher, optimum wet bulb approach will be from 8 degF [4.4 degC] to 12 degF [6.7 degC]. Other researchers (1) have gotten similar optimum wet bulb approaches. To keep a wet bulb approach of 8 degF [4.4 degC], a refrigeration plant will use significant additional condenser capacity, compared to a wet bulb approach of 12 degF [6.7 degC]. What is the major factor to determine optimum wet bulb approach? The major factor is condenser efficiency. This is defined as how much energy a condenser uses per unit of heat rejection. An evaporative condenser with axial fans and with integral(built-in) water pumps will use half the energy than an equal capacity condenser with centrifugal fans and with remote water pumps located in compressor room will use. To get optimum condensing pressure, a refrigeration plant with efficient axial fan condensers should use more condenser capacity because it is not energy intensive. An optimum wet bulb approach for these condensers would be 8 degF [4.4 degC]. On the other hand, to get optimum condensing pressure, a refrigeration plant with less efficient centrifugal fan condensers would use less condenser capacity because these are energy intensive. The optimum wet bulb approach would be 12 degF [6.7 degC]. Restated, the less efficient a condenser is, the greater the optimum wet bulb approach should be.

One operator has tried to optimize the wet bulb approach of his refrigeration plant and he got an optimum wet bulb approach of 27 degF [15 degC]. This approach is significantly greater than the 12 degF [6.7 degC] of an inefficient condenser. This indicated that the efficiency of the condensers is extremely bad. I don’t think that these condensers are so inefficient. It is likely that something was wrong with his testing. During optimization, to get a correct result, it is very important to have an in-depth understanding of the relationship between different factors that influence test variables. Just testing and comparing the number can give you a wrong result.

During summer operation, knowing the wet bulb approach is a useful feature for refrigeration plants with significant and frequent fluctuation of refrigeration load. To get the benefits of wet bulb approach control, a condenser should not be undersized. Typically, the capacity factor of the condensers should be 2 or greater. As mentioned above, a wet bulb approach from 8 degF [4.4degC] to 12 degF [6.7 degC] is optimum for the wet bulb temperature of 70 degF [21.1 degC] or higher. At lower wet bulb temperatures, properties of wet air will change and optimum wet bulb approach will be different.

Winter Operation

To get maximum energy savings, a refrigeration plant should operate at optimum set points for as long as possible. During winter operation, optimum condensing pressure can be as low as 50-60 psig [3.4 – 4.1 bars]. Most likely, plants cannot be operated at such a low condensing pressure, but this pressure should be kept as close as possible to the optimum. Reduction of the minimum allowable condensing pressure is a major energy saving measure. Very often this measure can provide up to 50% of total energy savings (2). Maximum effort should be given to reduce condensing pressure as low as possible. Unfortunately, the operators do not challenge traditional minimum condensing pressure of 110-120 psig [7.5 – 8.2 bars]. Sometimes lowering the condensing pressure by 10 psig [0.7 bars] will save more energy than compressor VFDs because lowering the condensing pressure will improve the efficiency of the whole refrigeration plant, but compressor VFDs will improve the efficiencies of only 1 or 2 compressors. A major misconception in industrial refrigeration is that a refrigeration plant should operate at condensing pressure above 100 psig [6.8 bars]. Operators do not challenge the traditional minimum condensing pressure of 100 psig [6.8 bars]. But I believe that majority of refrigeration plants can operate at condensing pressure below 100 psig [6.8 bars]. There are many imagined and real barriers to operating a plant at such a low pressure. However, every barrier has a solution. Very often there are several solutions and you will have the opportunity to choose a solution that is the best for your refrigeration plant. Let’s have a look at real barriers to operating the plants at low condensing pressure.

Barrier to Operating at Low Condensing Pressure #1: Hot gas defrosting

The typical minimum condensing pressure for hot gas defrosting is 110-120 psig [7.5 – 8.2 bars]. But at lower condensing pressures many plants have insufficient defrosting. PLCs can be programmed to increase condensing pressure for defrosting and reduce condensing pressure when defrosting is over. This approach can give us some energy savings, but the plants with many evaporator coils will always be on defrost and little energy will be saved. However, some plants use excessive condensing pressure to solve hot gas defrost problems. There can be many reasons (hot gas undersupply, hot gas oversupply, poor condensate draining and etc.) that can lead to poor hot gas defrosting. But instead of solving these reasons many companies just increase condensing pressure to improve defrosting. Typically this action will solve the issue, but efficiencies of the refrigeration plants will suffer.

Hot gas defrosting is a triple process, consisting of hot gas supply, ammonia condensation and ammonia condensate draining. It is easy to defrost an evaporator coil at a high condensing pressure, but it is not an easy task to do defrost at a low condensing pressure. To get adequate defrosting at a lower condensing pressure, the three parts of defrosting should be balanced. Misbalance of these parts is a major reason for poor hot gas defrosting. It is very important to determine which part is misbalanced. Many people believe that pressure drop in hot gas line at lower condensing pressure is a typical reason for poor hot gas defrosting. My experience has shown that this pressure drop is not significant. Typically, in the middle of defrosting, pressure drop in the hot gas line is less than 5 psig [0.34 bars]. It is easy to check my statement. Put the gages on the hot gas line and compare the readings.

Example
A cold storage facility has 24 evaporator coils. The main hot gas line was designed for simultaneous defrosting of 8 coils (one third of all coils). But during winter operation, only one evaporator coil per hour will be defrosted. The pressure drop in the hot gas main will not be significant because actual mass flow will be one eighth of designed mass flow.

Very often evaporator coils located in a penthouse have ice formation at the bottom of these coils. The reason for the ice build-up is cold air movement around the coil during defrosting. Operating evaporator fans create slight vacuum in the penthouse. When one coil goes on defrosting, other coils are running. Operating fans pull cold air through the penthouse opening, as well as through the defrosting coil. Melted frost will refreeze at the bottom of this coil. To solve this issue, all evaporators within a penthouse should be defrosted simultaneously or operating evaporators should idle when one of the evaporators in this penthouse goes on defrosting.

My experience has shown that very often hot gas is oversupplied to the defrosting coil. Why is it a bad idea to oversupply hot gas for the defrosting?

Safety. I have found that hot gas balancing valves of many evaporator coils are wide open. Sometimes evaporator coils are designed without hot gas balancing valves, but will instead have small additional hot gas pilot valves. These pilot valves will open first and pressure in the coils will gradually increase up to 30-40 psig [2-2.7 bars]. Then the main hot gas valves will open. Even using this approach, the main hot gas valve has a pressure difference of 80 psig [5.4 bars] or greater. This is a significant pressure difference and a wide open main hot gas valve will create high velocity of supplied hot gas. This high velocity is not an issue for a properly operated refrigeration plant. However, in some circumstances, the high velocity can lead to coil damage. For example, if for some reason, a coil was not pumped out properly, high velocity of hot gas can lead to “water hammer” in the evaporator coil and this coil can be damaged.

A properly designed hot gas main has a liquid drainer. But in real life, this drainer can be plugged by dirt, scale and etc. Liquid ammonia formed by the condensation of hot gas can collect in this line. As soon as the main hot gas valve opens, liquid ammonia from hot gas main will “fly” into a coil and can cause damage. A gradual supply of hot gas will help prevent these catastrophic events. In the worst case scenario, evaporator coil will not defrost properly, but at least it will not be damaged.

Efficiency. Many people believe that an increased supply of hot gas will lead to faster defrosting. I do not agree with this conclusion. The majority of low temperature evaporators are bottom feed overfed coils. Oversupply of hot gas will create a lot of ammonia condensate. This condensate does not easily drain from the coil because it will be pushed out through small orifices located at the inlet of each circuit. I have found that poor ammonia condensate draining is one of the main reasons for poor hot gas defrosting. Sometimes, an oversupply of hot gas will create a lot of blow by gas, which will go into the suction line creating a significant parasitic load. The safety and efficiency of the refrigeration plants can be improved by a balanced supply of hot gas into the evaporator coils during a defrosting cycle.

During winter operation, a majority of refrigeration plants in North America have optimum condensing pressure below 100 psig [6.8 bars]. To maximize energy savings, minimum allowable condensing pressure should be kept as low as possible. What is the minimum condensing pressure that can provide adequate hot gas defrosting? My experience has shown that minimum operating condensing pressures for hot gas defrosting are:
• Coolers, regular freezers – 70 psig [4.8 bars]
• Low temperature freezers, blast freezers, spiral freezers – 80 psig [5.4 bars]

These are not back pressure regulator settings, these are minimum condensing pressures. If your evaporators have defrosting at a higher condensing pressure, most likely you have room for improvement. It is easy to defrost at 120 psig [8.2 bars], but it is not easy to do defrosting at 70 psig [4.8 bars]. Defrosting at low condensing pressure requires precise adjustment of this process. Sometimes it takes additional 5-10 minutes to defrost at a low condensing pressure, but the overall process will be safer and more efficient than defrosting at a higher condensing pressure. To maintain optimum hot gas supply to the coils throughout the year, every refrigeration plant should have outlet pressure regulator to keep constant pressure in hot gas main regardless of the system’s condensing pressure. Otherwise, the hot gas supply would need be adjusted several times per year based on actual condensing pressure.

Barrier to Operating at Low Condensing Pressure #2: Liquid supply

Liquid supply to the evaporators can be a barrier to lower condensing pressure. Typically, operators do not challenge this issue. One “optimized” refrigeration plant had a minimum allowable condensing pressure of 110 psig [7.5 bars] because at any lower condensing pressure liquid ammonia from the high pressure receiver would be undersupplied to the evaporator at the far end of the plant. The suction pressure of this evaporator was 30 psig [2 bars]. The pressure difference that supplies liquid ammonia to this evaporator, is 110 – 30 = 80 psig [5.4 bars]. Is 80 psig [5.3 bars] not enough pressure to supply the liquid ammonia to the far end of the plant? A typical liquid ammonia pump has 30 – 40 psig [2 - 2.7 bars] head pressure and it supplies liquid ammonia to every corner of the plant. What is the reason for this phenomenon? A liquid ammonia pump supplies subcooled liquid which compensates for pressure drop in the liquid line and ensures that ammonia is delivered to the evaporators in a liquid state. When ammonia is delivered from a high pressure receiver (110 psig [7.5 bars]) to the evaporator (30 psig [2 bars]), a portion of the liquid can evaporate (flash) due to pressure drop because ammonia in the high pressure receiver is in a saturated state. A mixture of liquid and vapor can be delivered to the metering device and vapor will “choke” this device. The evaporator would be undersupplied with liquid ammonia. To solve this issue we have two choices:
• Subcool the liquid ammonia
• Increase size (Cv) of metering device

There are two ways to subcool liquid ammonia. A liquid ammonia pump can increase the pressure of the liquid and it will become subcooled. Another way to subcool is to reduce the temperature of saturated liquid ammonia. Very often the temperature can be reduced by 5-10 degF [2.8 - 5.6 degC] to solve the issue. Typically, a DX subcooler or a high pressure coil in an intermediate receiver can provide this subcooling.
Sometimes high temperature coolers and docks have DX evaporators with thermostatic expansion valves (TXVs) as metering devices. The refrigeration plants will operate at a higher than necessary condensing pressure to supply liquid to these TXVs. Operators were correctly told that at a lower condensing pressure, the capacity of these TXVs will be reduced. But do we need the maximum refrigeration capacity of these evaporators all year around? No. A major part of the refrigeration load of these coolers is heat transmission and infiltration from ambient air. Therefore during periods of cool weather, condensing pressure can be decreased and the capacity of DX evaporators will be reduced. But the refrigeration load will also be reduced. Thus demand and supply will be balanced and temperatures in refrigerated rooms will be kept at a desired level. Consequently, I think that very often the need for high condensing pressure to operate TXVs is an imagined barrier.

Some screw compressors have liquid injection oil cooling. Usually, liquid refrigerant from the high pressure receiver is injected into the compressor at an early stage of compression. The TXVs regulate the flow of liquid to maintain a desired oil temperature. Compressor manufacturers often specify a minimum condensing pressure of 120 – 130 psig [8.2 - 8.8 bars] to ensure adequate injection performance. To keep oil temperature at a desired level, demand and supply should be balanced. The demand is oil cooling heat rejection. The supply is a liquid refrigerant flow. Reduction of the condensing pressure will lead to a decrease of demand and supply, but these decreases will be at different rates. The liquid supply will decrease faster than heat rejection. At a certain lowered condensing pressure, an imbalance will happen and oil will overheat. Little can be done about the demand (oil cooling heat rejection), but liquid supply can be increased in several ways. These ways are:
• Increase size (Cv) of metering device. We cannot just increase the size of a traditional pressure/temperature controlled TXV because it can operate only in a narrow range of condensing pressures. A small TXV will operate properly at a high condensing pressure, but it will not operate properly at a low condensing pressure because it will underfeed the compressor. A large TXV will operate properly at a low condensing pressure, but will not operate properly at a high condensing pressure because it will overfeed the compressor. However, an electronic expansion valve can operate at a wide range of condensing pressures and can be a good replacement for a traditional pressure/temperature controlled TXV.
• Increase inlet pressure of metering device. This pressure can be increased by increasing condensing pressure, but this approach is not energy efficient. At a low condensing pressure, a liquid pump can be used to increase the pressure of a liquid refrigerant and the efficiency of the refrigeration plant will significantly improve.
• Reduce outlet pressure of metering device. This pressure can be reduced by actions such as: changing the location of the liquid injection port, reducing the suction pressure and unloading the screw compressor. However, suction pressure reduction and a compressor unloading will reduce compressor efficiency. However, lower condensing pressure will increase compressor efficiency. These opposite factors will “fight” each other. To choose the right action, total system performance should be evaluated.

Example.
A screw compressor has TXV for liquid injection oil cooling. This compressor was unloaded to 90% of nominal capacity. Due to this action, the refrigeration plant efficiency was reduced by 2%. At the same time condensing pressure was reduced. Due to lower condensing pressure the efficiency of this refrigeration plan has improved by 7%. Total system efficiency has increased by 5%. This is a step in the right direction.

Every refrigeration plant is unique and every refrigeration plant has the best solution for liquid injection oil cooling. Unfortunately, many companies use the least efficient solution of increased condensing pressure.

Barrier to Operating at Low Condensing Pressure #3: Oil carry-over

Many refrigeration plants operate at a higher condensing pressure due to a concern about oil carry-over from the compressors. Sometimes, operators will not try to lower the condensing pressure because they were told about possible oil carry-over. It is a good operating practice to keep track of oil charging into the compressors and oil draining from the oil pots. So, initially the operators should determine the severity of this carry-over. To reduce the condensing pressure, begin by selecting a reduction of 5 psig [0.34 bars] and monitor operation of the plant for a week or two. If there is no significant difference in oil carry-over, lower the condensing pressure by an additional 5 psig [0.34 bars] and monitor again. Sometimes, at a lowered condensing pressure, 2 gallons [7.6 liters] of oil need to be drained instead of 1 gallon [3.8 liters] and an annual cost of additional oil will be $1,000. However, the energy savings will be $20,000 due to a lower condensing pressure. It is a good trade off that can give you net savings of $19, 000.

A lower condensing pressure will reduce the density of a compressor’s discharge gas. At relatively constant mass flow, reduced discharge gas density will increase volume flow and thus the velocity through the oil separator. This increased velocity is the reason of oil carry-over. To reduce the discharge gas velocity, several actions can be done;
• Increase size of oil separator
• Reduce refrigerant mass flow

Refrigerant mass flow can be reduced by unloading a compressor and/or by reducing suction pressure. These two actions will reduce compressor efficiency. But at the same time, a lower condensing pressure will increase compressor efficiency. These factors are opposite. To choose the right action, an evaluation of the whole system performance should be done. For one plant it will be better to reduce suction pressure, for another plant it will be better to unload compressor. To prevent oil carry-over, many plants operate at a higher condensing pressure. This is a simple, but inefficient solution. I think that we have better options and the right one can be chosen.



Key Energy Saving Measure #2: Optimum suction pressure

It is common knowledge in the industry that raising the suction pressure improves compressor efficiency. Typical improvement might be 1-2% per 1 degF increase in suction temperature. However, to determine a system’s optimum suction pressure (temperature), the efficiency of the whole refrigeration plant should be evaluated.

Evaporators of many refrigerated rooms (freezers, coolers, docks and etc.) have single speed evaporator fans. Increased suction pressure will reduce the temperature difference between suction temperature and air temperature in a refrigerated room. To keep the required refrigeration capacity, the evaporator surface area should be increased and additional evaporators should be operated. To run additional evaporators, additional fan energy is required. This energy will be released in a refrigerated room and additional compressor energy is required to remove this parasitic refrigeration load. To estimate plant efficiency at a higher suction pressure, energy saved by compressors should be compared to energy spent by additional evaporator fans. Sometimes, a higher suction pressure will lead to an increase of total (compressors + evaporator fans) energy use. To keep the total energy use at a minimal level, the refrigeration plant should be operated at an optimum suction pressure or at an optimum temperature difference. A major factor of optimum suction pressure is the power of evaporator fans.

Example
There are 3 evaporator coils. The capacity of each coil is 20 TR [70 KW]. However, evaporator fans are different. Coil 1 has 6 BHP [4.5 KW] fans. Coil 2 has 10 BHP [7.5 KW] fans. Coil 3 has 15 BHP [11.2 KW] fans. When fan power increases, energy efficiency (power use per unit of refrigeration capacity) of evaporator coil will decrease. My research (3) has shown that these evaporators will have different optimum temperature differences or different optimum suction pressures. When the power of evaporator fans increases, the optimum temperature difference will increase as well. This is similar to condenser optimum wet bulb approach. The less efficient a unit (evaporator or condenser) is, the greater an optimum temperature difference or wet bulb approach will be. For the evaporator coils above, the range of optimum temperature differences (TD) will be as follows: coil 1optimum TD is 5 - 10 degF [2.8 – 5.6 degC], coil 2 optimum TD is 10 – 15 degF [5.6 – 8.3 degC], coil 3 optimum TD is 15 – 20 degF [8.3 – 11.1 degC]. To keep energy use of a refrigeration plant at a minimal level, it should be operated at an optimum suction pressure, regardless of refrigeration loads.



Key Energy Saving Measure #3: Optimized Hot Gas Defrosting

My experience has shown that defrosting of many refrigeration plants is not optimized. These plants have significant energy losses due to the wrong frequency of hot gas defrosting. To optimize this frequency, a criterion of optimization should be chosen. For cold storage facilities, the criterion should be the minimum total losses related to the frost and to hot gas defrosting. Frost on the evaporator coil reduces the capacity of this coil. To make up the lost capacity, an additional evaporator should be operated.

Example
Due to frost built up, a capacity of the evaporator will gradually decrease. At the end of the cooling cycle the capacity of this evaporator will be reduced from 100% to 90% of nominal capacity. Average coil capacity during one cooling cycle will be 95%. The evaporator capacity is 40 TR [140 KW]. The evaporator fan power is 20 BHP [15 KW]. Another evaporator should be operated 5% of the time to make up the lost capacity. This means that additional fan energy use will be 20 x 0.05 = 1 BHP/Hrs [0.75 KW/Hrs]. Typically, for every 1 BHP [0.75 KW] of fan power, an additional 0.5 BHP [0.37 KW] of compressor and condenser energy will be consumed to remove the heat of evaporator fans. The total penalty related to the frost will be 1.5 BHP/Hrs [1.12 KW/Hrs]. This penalty would be multiplied by the number of operating hours between defrosts.

The efficiency of hot gas defrosting is low and it is usually 20% or less (4). This means that for every 100 units of heat provided by hot gas, less than 20 units will be used for frost melting. Therefore, over 80 units must be removed by the refrigeration plant as a parasitic refrigeration load. Hot gas supply during defrosting usually doubles the amount of gas generated during a cooling mode. This means that after 35 minutes of hot gas defrosting, the evaporator coil, as well as part of refrigeration plant, will run for the next 1 hour in a cooling mode to remove the heat of defrosting.

Example
The evaporator mentioned above has a capacity of 40 TR [70 KW]. Operating time between defrosts is 12 Hrs. The efficiency of this refrigeration plant is 2.5 BHP/TR [0.53 KWe/KWr]. The refrigeration plant operating time required to remove the parasitic refrigeration load related to hot gas defrosting is 1 Hour. Losses related to one defrost is 40 x 2.5 = 100 BHP [75 KW]. Losses related to the frost will be 1.5 x 12 = 18 BHP [13.4 KW]. Total losses will be 100 + 18 = 118 BHP [88 KW]. Total operating hourly losses will be 118/12 = 9.83 BHP/Hrs [7.33 KW/Hrs]. Different frequencies of defrosting will lead to different total hourly losses. Optimum defrost frequency will give the minimum total hourly losses.



Key Energy Saving Measure #4: Optimized Compressor Sequence

Sometimes the operators try to change the compressor sequence and save energy by improving part load operation of their compressors. I think that the reasons for refrigeration load fluctuation should be determined first. Unfortunately, very often little attention has been given to the dynamic of the refrigeration plants operation. The majority of cold storage facilities have steady refrigeration loads. Why do evaporators of many cold storage facilities continuously cycle on and off? Why do compressors continuously load and unload, start and stop? This happens because evaporators typically operate independently. Air inlet temperature is a controlled variable. When air inlet temperature increases, an evaporator turns on. When air inlet temperature decreases, an evaporator turns off. However, we are interested in the refrigerated room temperature and the room temperature should be a controlled variable. To better control the room temperature and to minimize fluctuation of refrigeration load, the average air inlet temperature of operating evaporators would be monitored. When room temperature is satisfied, evaporators’ capacity should be reduced gradually by switching off the evaporators one by one with time delay. This approach will prevent significant fluctuation of the refrigeration load.

Example
A refrigeration room has 4 evaporators. The current refrigeration load is 70% of nominal load. If every evaporator operates independently, refrigeration plant load will fluctuate significantly. The number of operating evaporators can change from 0 to 4. However, all evaporators remove heat from the same room and the average room temperature should be a controlled variable. At 70% of nominal load, 2 evaporators should operate continuously and a third evaporator should run intermittently with time delay. Average air inlet temperature will be maintained at a desired level. Operating evaporators should be switched from time to time, to prevent hot spots in the refrigerated room.

The majority of intermediate pressure and low pressure receivers have solenoids and expansion valves for liquid make up. Solenoid valve cycling creates an artificial fluctuation of the refrigeration load due to fluctuation of flash gas generation. A modulating valve will be helpful to stabilize the flash gas refrigeration load. Slight changes in design and operation of a refrigeration plant can significantly improve the dynamic of a plant and part load operation and energy will be saved.



Key Energy Saving Measure #5: VFDs

VFDs can be helpful to recover losses related to part load operation (compressors, condensers, evaporators). But I think that the ability of VFDs to save energy is exaggerated. To estimate energy saved by VFDs one question should be answered. How great are the energy losses related to part load operation? If a plant is poorly designed and/or poorly operated, energy losses will be great and a lot of energy can be recovered through the use of VFDs and they will have a relatively short payback. If a plant has a good design and is operated well, energy losses will be minimal and little energy can be recovered. Payback for an investment in VFDs for this plant will be very long. Plant design and plant operation are two major factors to determine VFDs’ payback.

Example
A refrigeration plant has two single stage 500 BHP [373 KW] screw compressors. Suction pressure is 0 psig, - 28 degF SST [0 bars]. Condensing pressure is 120 – 150 psig, 75 degF-85 degF CT [8.2 – 10.2 bars]. A VFD for one compressor was installed. Why was this plant designed as a single stage? Considering the suction pressure and condensing pressures above, this plant should be designed as a two stage system or as a single stage system with an economizer. Why are these compressors of equal capacity? A screw compressor is slightly inefficient when it loaded at 50 – 100% and it is very inefficient when it loaded lower than 50%. It is often effective to choose different compressors such that part load losses can be reduced. It can be 3 compressors (200 BHP [149 KW], 300 BHP [224 KW], 500 BHP [373 KW]) or it can be 2 compressors (400 BHP [298 KW], 600 BHP [447 KW]). Why is the minimum condensing pressure 120 psig [8.2 bars]? Due to high pressure ratio, part load losses are greater at high condensing pressures than part load losses at lower condensing pressures. Decrease of the minimum condensing pressure from 120 psig [8.2 bars] to 90 psig [6.1 bars] will reduce compressor energy use of mentioned refrigeration plant (2 compressors) by 15% and losses related to part load operation will be reduced as well. This energy saving is comparable to energy savings that can be achieved by installing one 500 BHP [373 KW] compressor VFD operated at refrigeration load of 40%. This means that an efficiently operated refrigeration plant without VFD installed on a compressor will have the same efficiency as typically operated plant with a VFD installed on the compressor. The cost of a 500 BHP [373 KW] VFD (more than $100,000) can be saved. However, maximum energy savings would be achieved by using compressor VFD and by operating a refrigeration plant at low condensing pressure.

Condenser fan VFDs can save up to 8% of a refrigeration plant’s energy use. But typically, the energy savings will be significantly lower because the operating time at favorable energy saving conditions will be short. During summer operation, the majority of refrigeration plants run condensers at full capacity. Very little energy can be saved during summer time. During winter operation, many condensers use little energy. This means that energy losses of part load operation are minimal and little energy can be recovered. During spring and fall some refrigeration plants can get good energy savings from condenser fan VFDs, especially plants with centrifugal condenser fans. To get maximum savings from condenser fan VFDs, condensers should not be undersized and capacity factor of these condensers should be greater than 2.

Example
A refrigeration plant has relatively steady refrigeration load all year around. Condenser fan VFDs were installed for this plant. At different seasons, these VFDs provide different energy savings. During the summer, there is no energy savings achieved because all of the condenser fans operated at full speed to provide maximum condensers’ capacity. During fall and spring operation, the average energy savings were 5% of the whole plant energy use. During winter operation, these energy savings were 2% due to cold weather, little condensers’ capacity and condensers’ energy were used. Average condenser fan VFDs energy savings were (5+5+2)/4 = 3% of the whole plant energy use. Based on these energy savings, a potential payback for condenser fan VFDs can be estimated.

Evaporator fan VFDs can be useful to save energy, especially for evaporator coils located in penthouses. Electrical motors of these evaporator fans, typically, are oversized. Many designers are over focused on air throw and little attention is paid to natural air convection. Typically, penthouses are located in the middle of a cold room and the warmest air located under the ceiling is pulled evenly from each side of the room. This factor helps prevent hot spots in a refrigerated room with penthouses and many evaporator fans can operate successfully at a low speed. Ceiling hung evaporator fans should have a good throw to push warm air from the opposite upper corner of a cold room. Modern cold rooms with penthouses are typically high-rise construction. During the cooling mode, the room height will help to natural air convection and the warmest air will always enter the evaporators.



Key Energy Saving Measure #6: Remote Tuning Up

To optimize operation of every refrigeration plant, two major steps should be done. Optimum set points and optimum operating strategies should be determined and implemented. To fulfill these steps specific knowledge and experience are required. I found that using this knowledge and experience, the operation of many refrigeration plants can be optimized remotely. This means that an expert in system optimization does not necessarily need to visit the plant. Typically, an end user provides information about the refrigeration plant design and operation. Based on this information an expert can give advice to the operator about set points and operating strategies as well as advice about the implementation of these points and strategies.

Example.
A refrigeration plant operates at a condensing pressure of 120 psig [8.2 bars]. Information about plant design and operation was provided to an expert by an end user. The expert has reviewed this information and has determined that the optimum condensing pressure for this plant at current ambient conditions is 90 psig [6.1 bars]. He suggests the operator reduce the condensing pressure, monitor hot gas defrosting, liquid supply to the intermediate receiver and oil consumption. Condensing pressure was reduced 5 psig [0.34 bars] per step and the operation of the plant was monitored. At a condensing pressure of 110 psig [7.5 bars], the operator noticed that the bottom of evaporator coils did not defrost properly. The expert suggested how to readjust the hot gas supply to the operator. Readjustment of the hot gas supply was done and condensing pressure was gradually lowered to 100 psig [6.8 bars]. At this condensing pressure liquid supply to the intermediate receiver was readjusted and the condensing pressure was lowered to 90 psig [6.1 bars]. These 6 steps of tune up were done within 3 weeks.

Tuning of a refrigeration plant should be done gradually to avoid shocks to the operators and to the plant. Sometimes people operate refrigeration plants at certain set points for years. It can be a shock to them and to the plant, if set points are changed significantly in one step. Gradual adjustment is preferred because operators will have time for reaction if any concerns arise. Using this approach the expert does not require a visit to the plant. This means that there is no upfront investment for the end user. Typically, the initial part of the tuning up is free. It gives an opportunity for the end user to test this approach without any investment. I think that many industrial ammonia refrigeration plants around the globe can benefit from remote tuning up. Therefore, to maximize energy savings end users should be open to this new approach.



Cost effective energy savings

I think that today every company is interested in good investment in energy efficiency. What is the goal of these investments? Everybody is interested in investing less money and in getting a better return on their investment. Optimization of a refrigeration plant operation can give us the best return with minimum investment. Initial optimization can be done by the plant operators. Management should find a way to involve their operators in active energy saving process. However, these are low hanging fruits. What is the next step that should be done? Two options are available. Option 1: Final optimization of the refrigeration plant operation. Option 2: Capital investments in energy saving equipment. Typically, capital investments have several years payback. Generous government and/or utility incentives can reduce this payback to an acceptable 2-3 years. This is the best payback you can get on the capital investments. But consider the payback on investments in final optimization of the refrigeration plant operation. Typically, this payback is 2-3 months. This is 12 times (or 1200%) shorter payback than the payback on capital investments. In an investment industry, a higher return will lead to a higher risk of investment. But this is not an issue for investments in optimization. Capital investments in energy saving equipment have a higher risk because to test this equipment you have to buy it first. If you are not satisfied, you cannot return it. The initial part of the final optimization, typically, is free. It means that testing of this approach can be done without investing a single dollar. Would you prefer to invest your own money at 120% return, instead of 10% return and without any risk? Unfortunately, this option is not available on financial market, but it is available in industrial refrigeration. Tuning up of a refrigeration plant operation can give us this fantastic opportunity.



Conclusion

Optimization of a refrigeration plant operation is a corner stone of energy saving process in industrial refrigeration. Balanced plant operation does not require significant financial investment, but it does require a high level of expertise to analyze the operation of the refrigeration plant as a whole system. Fortunately, our industry has knowledge and experience to significantly improve the operation of many refrigeration plants. To maximize efficiencies of their refrigeration plants, end users should be open minded to new approaches in energy savings. Engineering operation of the refrigeration plants should be the ultimate goal of every company. Optimization of the refrigeration plants operation will give our industry a unique opportunity to significantly improve efficiencies of many refrigeration plants without significant financial investments and maximum energy savings will be achieved if the right approach to optimization is chosen.


References:
1. Reindl, D.T., Jekel, T.B., Elleson, J.S. Industrial Refrigeration Energy Efficiency Guidebook. Industrial Refrigeration Consortium – University of Wisconsin-Madison. Madison, WI. 2005
2. Russell U.K., “The Process of Optimizing and Fine Tuning Refrigeration Systems to Improve Efficiency and Lower Operating Costs”, 2005 IIAR Ammonia Refrigeration Conference Acapulco, Mexico, International Institute of Ammonia Refrigeration, Arlington, VA, (2005)
3. Khoudiachov, S. “Ideas on Energy Savings” IIAR Magazine “Condenser “November 2008.
4. Stoecker, W.F., Industrial Refrigeration Handbook, McGraw Hill publishers, 1998



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