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Operating Ammonia Refrigeration Systems At Peak Efficiency

S.Khoudiachov
S. K. Energy Consulting
69 Laurier Ave. Richmond Hill
Ontario. L4E4P5. Canada
Fax: (905) 773-7947


Abstract


There are several ways to improve energy efficiency of ammonia refrigeration systems. To save energy, many companies invest in additional condensers, sophisticated PLC’s, and VFD’s. Unfortunately, optimization of the refrigeration plant operation is often overlooked. However, this optimization is the most cost-effective approach to save energy in ammonia industrial refrigeration. By improving the operation of ammonia systems, companies can get 10 – 20 times better return on investments compared to typical investments in energy saving equipment. This technical paper will review practical implementation of the optimized operation. Many existing refrigeration systems can save 10- 20% of the energy use by optimizing condensing pressure, suction pressure, hot gas defrosting, etc. To achieve these energy savings, limited investments are required. However, operators should work as a team with energy efficiency experts to minimize energy use of their refrigeration systems.

1. INTRODUCTION


Today many companies focus on sustainability and energy efficiency. There are several ways to save energy in ammonia industrial refrigeration. Some companies invest in variable frequency drives (VFD’s) and in additional condenser capacity; other companies improve design of their refrigeration plants and buy sophisticated PLCs. Definitely, these approaches save energy, but they are very expensive. However, during current global economic uncertainties companies pay more attention to cost effective energy savings. Optimization of the refrigeration plant operation is the most cost-effective way to save energy in our industry. This optimization doesn’t require significant investments, but significant energy savings can be achieved. To maximize energy efficiency of any refrigeration plant, it should be operated at optimum set points and at optimum operating strategies all year around. However, operators should be able to determine these set points and operating strategies, and they should be able to implement these set points and strategies.

2. SUMMER OPTIMUM CONDENSING PRESSURE


To maximize the energy efficiency of a refrigeration plant, it should be operated at optimum condensing pressure as long as possible. At this pressure total energy use of high side (compressors + condensers) of the refrigeration plant is minimal. How to determine the optimum condensing pressure? How to implement the required condensing pressure?
During summer operation, condensing pressure of the ammonia refrigeration plant with evaporative condensers should “float” up and down based on fluctuation of the wet bulb temperature of the ambient air. To provide this “floating”, modern PLCs have wet bulb approach control feature. This approach is the temperature difference between condensing temperature and wet bulb temperature of ambient air and it keeps the balance between compressor and condenser capacities. At optimum wet bulb approach, total energy use of the high side of the refrigeration plant is minimal. Typically, summer optimum wet bulb approach ranges from 4.4 degC to 6.7 degC (1). What is the major factor to determine optimum wet bulb approach? The major factor is condenser energy efficiency. This is defined as how much energy a condenser uses per unit of heat rejection. Evaporative condensers with axial fans use less energy and their optimum range is 4.4 degC to 5.6 degC. Condensers with centrifugal fans use more energy and their optimum range is 5.6 degC to 6.7 degC. Every refrigeration plant has its own optimum summer wet bulb approach and it should be kept to maintain the right balance between compressor and condenser capacities. However, many people are over focused on “floating” condensing pressure. Definitely, this “floating” saves energy, but these savings are not significant.
Example.
Refrigeration plant has 3 equal screw compressors. However, at current refrigeration load 2 compressors are operated. Total compressor energy use is 347.5 kW. Evaporative condenser has water pump of 14.9 kW and fans of 33.6 kW.
Initially, this plant was operated at fixed condensing pressure of 1046 kPa. To keep this pressure, 11.2 kW of fan power was required. Recently, operation of this plant has been changed to “floating” condensing pressure. Currently, condensing pressure “floats” from 827 kPa to 862 kPa. Action of switching operation from fixed to floating condensing pressure can be divided in 2 steps.
Step 1. Condensing pressure was lowered from fixed (P1) 1046 kPa to optimum (P2) 862 kPa. This step saved 37.3 kW of compressor power, but added 11.2 kW of fan power. 37.3 – 11.2 = 26.1 kW of energy were saved.
Step 2. Lowering condensing pressure from 862 kPa to 827 kPa will save 7.5 kW of compressor power, but 11.2 kW of fan power will be added. Total energy use will increase by 11.2 – 7.5 = 3.7 kW. “Floating” condensing pressure (P2) will save 3.7 kW, compare to operating plant at minimum fixed (P3) condensing pressure of 827 kPa. As we can see, 26.2 kW of energy saved by lowering condensing pressure from P1 to P2 is significant energy saving step and it is 7 times greater than 3.7 kW of energy saved by “floating” condensing pressure.



When can we “float” condensing pressure? Typically, condensing pressure can be floated when wet bulb temperature of ambient air is between 10 degC and 21 degC. If wet bulb temperature of ambient air is below 10 degC, majority of the refrigeration plants operate at minimum allowable condensing pressure and there is no “floating”. Usually, if wet bulb temperature is above 21 degC, condensers operate at full capacity and there is no “floating” as well. Unfortunately, real “floating” of the condensing pressure can be applied for a limited period of time.
To achieve the benefits of “floating” condensing pressure, a refrigeration plant should have a sophisticated PLC with wet bulb approach feature. Unfortunately, many refrigeration plants do not have wet bulb approach feature. How to keep optimum condensing pressure for these plants?
Two options are available:
1. Wet bulb approach feature is helpful to maintain certain balance between compressor and condenser capacities. However, optimum balance between the compressor and condenser capacities can be kept without wet bulb approach. This option works well for plants with steady refrigeration load. Typically, cold storages have relatively steady load and optimum balance can be kept. How can we do that? First, optimum balance between compressor and condenser capacities should be determined. Then, this balance should be implemented.
Example.
Current refrigeration load of the plant is 100%. Four compressors of equal capacities should be operated to handle this load. This plant has 3 equal condensers. However, total condenser capacity is oversized. It was determined that optimum condensing pressure will be achieved by operating 2 condensers. This plant has optimum balance when 4 compressors and 2 condensers are operated.
When refrigeration load is 50%, 2 compressors will be operated. To keep the balance between compressor and condenser capacities, 1 condenser should be operated as well.
2. Loads of many refrigeration plants have significant fluctuation during production time. Different operating strategy should be chosen for these plants. Typically, wet bulb temperature changes gradually. It increases during the day and decreases during the night. To implement optimum condensing pressure, set points of the condensing pressure should be changed 3 - 4 times per day. Based on wet bulb temperatures of ambient air, a table of optimum condensing pressures can be created. From time to time, the operators should check wet bulb temperature of ambient air and by using the mentioned table they will be able to determine and to implement optimum condensing pressure.

3. WINTER CONDENSING PRESSURE


In industrial refrigeration, energy can be saved when opportunity exists. Typically, during summer operation, a refrigeration plant operates at full capacity and it is challenging to save energy at this time of the year. However, during winter operation many energy saving opportunities arise. Mother Nature gives us these opportunities, but very often we don’t use them to their full extent.
During winter operation, optimum condensing pressure can be as low as 350 – 400 kPa. Is it
possible to operate ammonia refrigeration plant at such low pressure? Most likely not, but any effort should be done to reduce operating condensing pressure to keep it 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). A major misconception in industrial refrigeration is that a refrigeration plant should operate at condensing pressure above 700 kPa. Operators do not challenge the traditional minimum condensing pressure of 700 kPa. But I believe that the majority of refrigeration plants can operate at condensing pressure below 700 kPa. However, there are many imagined and real barriers to operate plants at such a low pressure, but every barrier has a solution. Very often there are several solutions and the best one, for the particular refrigeration plant, can be chosen. Let’s have a look at real barriers to operate the plants at low condensing pressure.

3.1 Hot gas defrosting
The typical minimum condensing pressure for the hot gas defrosting is 760 – 830 kPa, because at lower condensing pressures many refrigeration 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 refrigeration plants with many evaporators will always be on defrost and little energy will be saved by reducing condensing pressure.
Hot gas defrosting is a triple process, consisting of hot gas supply, ammonia condensation and ammonia condensate draining. To get adequate defrosting at a low 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 where the misbalance is. Many people in our industry believe that pressure drop in hot gas line is a typical reason for poor hot gas defrosting at low condensing pressure. However, I think that this is a misconception. Typically, hot gas main is designed for simultaneous defrosting of 25 – 30% of the evaporators. During periods of cool weather, not all evaporators should be operated and operating evaporators should not be defrosted as often as during summer operation. Typically, 4 – 6% of the evaporators should be defrosted simultaneously. Actual mass flow in hot gas main will be at least 5 – 6 times lower than designed mass flow and the pressure drop in this line will not be significant.
The majority of low temperature evaporators are bottom feed overfed coils. To defrost them faster, operators increase hot gas supply to the defrosting coil. Oversupply of hot gas creates a lot of ammonia condensate. This condensate can not be easily drained from the evaporator because it should be pushed out through small orifices located at the inlet of each circuit. Bottom of the evaporator will not defrost unless it is free of liquid ammonia. Inadequate ammonia condensate draining is one of the main reasons of poor hot gas defrosting. Adequate condensate draining can be achieved by keeping pressure difference between defrosting coil and defrost back pressure regulator of 70 – 100 kPa for cooler and 170 – 210 kPa for freezers. Required pressure difference can be achieved by adjusting hot gas supply and defrost back pressure regulator. During defrosting cycle, energy efficiency of the refrigeration plants can be improved by a balanced supply of hot gas into the evaporator coils as well as adequate condensate draining.

3.2 Liquid supply
Some refrigeration plants use condensing pressure to supply liquid ammonia to the evaporators. Very often operators claim that they can not lower condensing pressure below 750 – 800 kPa, because at any lower condensing pressure liquid ammonia from the high pressure receiver would be undersupplied to the evaporators at the far end of the plant.
Liquid ammonia in the high pressure receiver is in a saturated state. When ammonia is delivered from a high pressure receiver to the evaporator, a portion of the liquid will evaporate (flash) due to pressure drop in liquid supply line. A mixture of liquid and vapor will 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 liquid ammonia or increase size (Cv) of the metering device.
There are two ways to subcool liquid ammonia. A liquid ammonia pump can increase the pressure of the liquid ammonia 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 2 – 3 degC to solve the issue. Typically, a DX subcooler or a high pressure coil in an intermediate receiver can provide this subcooling.
Size of the metering device can be increased as well. Two solenoids can be used in parallel to operate in wide range of the condensing pressures. The size of a traditional pressure/temperature controlled TXV cannot be just increased because it can only operate in a narrow range of condensing pressures. 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.

3.3 Oil carry-over
Many refrigeration plants operate at a higher condensing pressure due to a concern about oil carry-over from the screw 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 35 kPa 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 35 kPa and monitor again. At lowered condensing pressure oil carry-over can slightly increase and the annual cost of oil required to charge a refrigeration plant can increase by $1,000. However, the energy savings can be $10,000 due to a lower condensing pressure. It is a good trade off that can give us net savings of $9, 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, two actions can be done: increase size of oil separator or 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.
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.

4. OPTIMUM SUCTION PRESSURE


It is common knowledge in the industry that raising the suction pressure improves compressor energy efficiency. Typical improvement might be 3 - 4% per 1 degC increase in suction temperature. Based on this information many cold storages use the following operating strategy. If all room temperatures are satisfied and evaporators are in a low load mode, the suction pressure usually increases to the maximum until some limiting temperatures are approached; thus, the efficiency of the compressors will improve. However, to minimize energy use of the refrigeration plant, the energy efficiency of the whole refrigeration system should be evaluated and a system’s optimum suction pressure/temperature should be determined.
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.
To determine optimum suction temperature/pressure, tables 1 & 2 were calculated for 2 refrigeration plants with evaporators that have capacity of 70.34 kWr at temperature difference of 5.6 degC (kWr is kilowatt of refrigeration effect). Evaporators of first plant have fan power of 4.47 kWe per unit and evaporators of second plant have fan power of 11.2 kWe per unit (kWe is kilowatt of electrical energy). Temperature in the refrigerated rooms is -18 degC. Figure 3 & 4, which were created based on Tables 1 & 2, show how several variables change with change of the temperature difference between suction temperature and air temperature in the refrigerated room. These variables are compressor energy efficiency, evaporator energy efficiency, system energy efficiency (energy use of compressors and evaporators per unit of refrigeration) and net system energy efficiency (energy use of compressors and evaporators per unit of net refrigeration). Net refrigeration is refrigeration capacity of the evaporator minus heat produced by evaporator fans

Table 1: Optimum Suction temperature for Evaporator Fan Power of 4.47 kW

Temperature
Difference,
deg C
Compressor
Efficiency,
kWe/kWr
Evaporator
Efficiency,
kWe/kWr
System
Efficiency,
kWe/kWr
Net System
Efficiency,
kWe/kWr
2.780.26220.12720.38940.4463
5.560.2860.06360.34960.3733
8.330.31140.04240.35380.3702
11.110.3390.03180.37080.3829
13.890.36930.02540.39470.4051





Figure 3: Optimum Suction temperature for Evaporator Fan Power of 4.47 kW




Table 2: Optimum Suction temperature for Evaporator Fan Power of 11.2 kW

Temperature
Difference,
deg C
Compressor
Efficiency,
kWe/kWr
Evaporator
Efficiency,
kWe/kWr
System
Efficiency,
kWe/kWr
Net System
Efficiency,
kWe/kWr
2.780.26220.3180.58020.8544
5.560.1590.2860.4450.5292
8.330.31140.1060.41740.467
11.110.3390.07950.41850.4545
13.890.36930.06360.43290.4624






Figure 4: Optimum Suction temperature for Evaporator Fan Power of 11.2 kW



Figure 3 shows that optimum temperature difference for the plant with 4.47 kWe evaporator fans is around 7 degC. Reduction of temperature difference to 2.78 degC will increase energy use of this refrigeration plant by 21%. Figure 4 shows that optimum temperature difference for the plant with 11.2 kWe evaporator fans is 11.11 degC. Reduction of temperature difference to 2.78 degC will increase energy use of the refrigeration plant by 88%. These numbers show that every refrigeration plant should be operated at optimum suction pressure instead of the highest possible suction pressure, otherwise energy efficiency of this plant will suffer.
Real life optimum temperature difference will be slightly greater than calculated by the follow factors: suction pressure losses, frost on the evaporators, actual fan power increase due to density of cold air and static pressure penalty.

5. OPTIMUM HOT GAS DEFROSTING


Many refrigeration plants have significant energy losses due to hot gas defrosts of the evaporators. Wrong frequency of these defrosts is a major reason of these losses. To optimize this frequency, a criterion of optimization should be chosen. For cold storage facilities, the criterion should be the minimum hourly 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 and additional fan energy will be used. This is the first part of total losses. The efficiency of hot gas defrosting is low and it is usually less than 20% (3). Assume that efficiency of the defrost is 10%. It means that 10% of supplied heat was used for frost melting and 90% of supplied heat was released in the refrigerated room as parasitic refrigeration load. This defrost is the second part of total losses. To improve energy efficiency of the refrigeration plant, total losses related to the frost and to the defrosting should be minimal.
Example.
Assume that after 20 Hrs of operation, capacity of the evaporator have been reduced to 90% of nominal capacity. It means that average hourly evaporator capacity was 95%. Additional evaporator should be operated for 1 Hr or 5% of the operating time, to make up capacity that was lost due to frost formation. Nominal capacity of this evaporator is 80 kWr. Fan power of this evaporator is 8 kWe. Hourly losses related to frost will be (1 x 8)/20 = 0.4 kWe/Hrs. After 20 Hrs of operation, hot gas defrosting was initiated. Full defrost time is 1 Hr, includes pump out, defrost, pump down and time delay. Hot gas was supplied for 0.5 Hr. Efficiency of this hot gas defrosting is 10%. Hot gas supply during defrosting usually equals to 1.5 the amount of gas generated during a cooling mode. Amount of heat released into refrigerated room can be calculated as 80 x 1.5 x 0.5 x 0.9 = 54 kWr. Assume that energy efficiency of this refrigeration plant is 0.4 kWe/kWr. Total losses of electrical energy related to hot gas defrosting will be 54 x 0.4 = 21.6 kWe. Hourly losses related to hot gas defrosting will be 21.6/20 = 1.08 kWe. Total hourly losses related to the frost and to the hot gas defrosting will be 0.4 + 1.08 = 1.48 KWe.
At optimum hot gas defrosting frequency, total hourly losses will be minimal. Two major variables influence optimum frequency: rate of frost formation and evaporator fan power. Two tables were created for 3 different fan powers (4 kWe, 8 kWe and 12 kWe) and two different rates of frost formation. At first frost formation rate (summer), evaporator capacity was reduced to 90% in 10 Hrs. At second frost formation rate (winter), evaporator capacity was reduced to 90% in 20 Hrs.


Table 3: Summer total hourly losses

 Evaporator Capacity % (Operating hrs)
 95(5)90(10)80(20)70(30)
Evap.
Fan
Power
kWe
44.422.361.481.32
84.522.561.881.92
124.622.762.282.52



Table 4: Winter total hourly losses

 Evaporator Capacity % (Operating hrs)
 95(5)90(10)80(20)70(30)
Evap.
Fan
Power
kWe
42.261.280.940.96
82.361.481.341.56
122.461.781.742.16


As we can see from Tables 3 & 4, optimum operating time before defrosting can vary from 20 to 50 Hrs. However, many refrigeration plants have defrost 4 times per day all year around. This wrong frequency can lead to significant energy losses.
To improve energy efficiency of a refrigeration plant, frequencies of hot gas defrosting should be changed 2 -3 times per year based on rate of frost formation.

6. CONCLUSION


Energy efficiency of many refrigeration plants can be significantly improved by implementing major energy savings measures, which were presented in this paper. Initial steps of this implementation can be done by the operators. However, to achieve optimum operation, a high level of specific knowledge and experience is required. Fortunately, our industry has this knowledge and experience to achieve maximum efficiency of many refrigeration plants. However, an end user should find the right expert to get help to optimize operation of his refrigeration plants. Optimized operation gives our industry a unique opportunity to significantly improve energy efficiencies of many refrigeration plants with minimal 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., 2005, Industrial Refrigeration Energy Efficiency Guidebook Industrial Refrigeration Consortium – University of Wisconsin-Madison. Madison, WI, 210 p.
2. Russell U.K. 2005, “The Process of Optimizing and Fine Tuning Refrigeration Systems to Improve Efficiency and Lower Operating Costs”, IIAR Ammonia Refrigeration Conference Acapulco, Mexico, IIAR, Arlington, 23-41
3. Stoecker, W.F. 1998, Industrial Refrigeration Handbook, McGraw Hill publishers, 782 p.


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