MCQUAY Long Piping Application Manual First Edition November 2005

Transcription

1 Long Piping

2 MCQUAY Long Piping Applica pplication Manual First Edition November 2005

3 Introduction Introduction... Introduction-1 Objective... Introduction-2 Section 1: Long Piping Limitation tion 1.1 Capacity Loss Oil Return Problem Compressor Failure High Installation Cost Unit Orientation Piping Length Limit Section 2: Precautions on Long Piping Installations tions 2.1 Additional Oil Additional Refrigerant Oil Trap Suction Accumulator Oil Separator Pipe Sizing Crankcase Heater Pump Down Cycle Minimize Bends Minimize Excessive Height Section 3: Capacity Effect 3.1 Pressure Drop Charts Section 4: Piping Installation tion 4.1 Pipe Material Pipe Insulation

4 4.3 Horizontal Runs Vertical Runs Pipe Bends Vibration and Noise in Piping Appendix A : Common Compressor Failur ailure A.1 Liquid Flood Back... A-2 A.2 Flooded Starts... A-3 A.3 Liquid Slugging... A-4 A.4 Loss of Lubrication... A-5 A.5 Overheating... A-7 A.6 Contamination... A-9 A.6 Refrigerant Migration... A-12 Appendix B : Suction Accumula ulator B.1 Applications... B-3 B.2 Installation... B-4 B.3 Sizing and Selecting an Accumulator... B-4 Appendix C : Oil Separ parator C.1 Introduction... C-1 C.2 Function... C-3 C.3 Installation... C-3 C.3 Maintenance... C-4 Appendix D : Useful Table and Charts Copyright 2005 by MCQUAY International. All rights reserved. This publication is strictly confidential and is meant for DISTRIBUTORS of MCQUAY International only. No part of this publication may be reproduced or distributed in any form or by any means, or stored in a database or retrieval system, without the prior written permission of MCQUAY International. The data and suggestions in this manual are believed current and accurate at the time of publication, but they are not a substitute for trained, experienced professional service. Individual applications and site variations can significantly affect the results and effectiveness of any information. The reader must satisfy him/herself regarding the applicability of any article and seek professional evaluation of all materials. MCQUAY International disclaim any responsibility for actions based on this manual.

5 Introduction Introduction Split type air-conditioner units consist of an evaporator unit and condensing unit which are both joined together by two lengths of copper piping. Generally, one of them will be in the indoor room while the other will be located outdoors. The performance specifications of the air-conditioner unit have been given based upon a specified pipe length. Deviations from this standard length will cause variations to the unit performance. There is also a maximum pipe length allowed for these units, whereby if exceeded, the unit may not give reliable performance. Most manufacturers will publish these longest distances in both the vertical and horizontal directions where their evaporator and condenser can be installed apart. Most of the time, installers are not aware of these limitations. As a result, they encounter problems when the units are not performing as specified. It is important that during installation, these pipe length limits are not exceeded. It is recommended that the pipe lengths be as close to the standard lengths as possible. The relative location of both evaporator and condensing units are also very crucial to ensure an effective and reliable system. However, this will not be so easily achievable in practice. The building architectural and structural design may not allow such straight forward installations. It is very common for these pipe lengths to be longer than the standard lengths, if not exceeding them. Therefore, it is vital to understand what the failure mechanisms of the air-conditioner unit are when this happens. With this in mind, we can then take the necessary precautions to prevent damage to the units. Introduction - 1

6 Objective The purpose of this Application Manual is to give guidelines into long piping installations of split type air conditioner units. It gives recommendations on the necessary precautions and modifications which are needed to be carried out to maintain the life span of the system. Information is also given on some of the common compressor failures encountered with such installations and what are the counter-measures which can be used to prevent them. The topics covered in this manual are as follows: a) Long piping limitations. b) Precautions on installation. c) Changes in capacity performance due to long piping. d) Practical piping installation. Introduction - 2

7 Section 1 Long Piping Limitations There are a few points that an installer and contractor need to consider when dealing with long piping installations. These points are: a) Capacity loss of the system. b) Problem of oil return to the compressor. c) Compressor failure. d) High installation cost. 1.1 Capacity Loss Capacity loss is due to pressure loss which is caused by friction in pipe and elevation. Consider the Darcy- Weisbach equation: p = f (L/D) (ρv 2 /2)... (1.1) where, p = Pressure drop, Pa. f = Friction factor, dimensionless. L = Length of pipe, m. D = Internal diameter of pipe, m. ρ = Fluid density at mean temperature, kg/m 3. V = Average velocity, m/s. Therefore: p L and V 2 p 1/D This equation can also be expressed in the form of specific energy: h = p/ρg = f (L/D) (V 2 /2g)... (1.2) And: where, h = Energy loss, m. g = Gravity acceleration, m/s 2. h L and V 2 h 1/D 1-1

8 The remaining terms are the same as mentioned in equation (1.1). By considering equation (1.1), it is known that L is directly proportional to p. When L increases (with other terms remaining constant); p will increase proportionally as well. In other words, when the piping length increases, the pressure drop encountered will be increasing proportionally as well. Again, considering equation (1.1), it is known that D is indirectly proportional to p. When D decreases (other terms remaining constant), p will also increase proportionally as well. Therefore, when the piping diameter decreases, the pressure drop encountered will be increasing proportionally as well. The pressure drop and capacity drop are directly related to each other. If the pressure drop decreases, the capacity drop will be also do the same. If the piping length is long, the flow will encounter a higher pressure drop. Therefore, the capacity for longer piping will decrease as well. Same for decreasing pipe diameter, the pressure drop will be large and hence, the capacity will drop as well. Why will the pressure drop cause a reduction in capacity? Basically, there are three reasons: a) Suction line pressure drop due to friction loss will force the compressor to operate at a lower suction pressure with a resultant reduction of refrigerant mass flow. b) Pressure drop due to friction loss in discharge lines causes the compressor to operate at a higher pressure resulting in reduced refrigerant mass flow and increased power consumption. c) Liquid line pressure drop due to friction loss and liquid static head may cause flash gas. This flash gas will reduce the performance of the expansion device as the liquid column cannot be maintained. 1-2

9 1.2 Oil Retur eturn n Prob oblem The functions of oil in refrigeration systems are: a) Minimize mechanical wear b) Reduce friction c) Lubricate moving parts d) Seal clearances e) Deaden noise, and f) Assist to transfer heat. In the compressor, oil and refrigerant will mix continuously. Refrigeration oils are soluble in liquid refrigerant and mix completely at normal room temperatures. Since oil must pass through the compressor cylinders to provide lubrication, a small amount of oil is always in circulation with the refrigerant. Oil and refrigerant vapour do not mix readily and the oil can be properly circulated through the system if gas velocities are high enough to sweep the oil along. If refrigerant velocities are not sufficiently high, oil will tend to lie at the bottom of the evaporator tubing which, decreases heat transfer efficiency and possibly causing a shortage of oil in the compressor. The problem arises when the piping is longer than the standard testing length of 7.6m. More oil will tend to be trapped along the longer piping and only a lesser amount will be returned to the compressor. This damages the compressor internal moving parts due to lack of lubrication. As evaporative temperatures are lowered, this problem becomes more critical since the viscosity of the oil increases with a decrease in temperature. Since the longer piping will cause a lower evaporating pressure due to the higher pressure drop, the oil becomes more viscous and more difficult to be swept along with the refrigerant. For these reasons, a proper design of tubing is essential for satisfactory oil return in a refrigeration cycle. One of the basic characteristics of a refrigerant and oil mixture in a sealed system is the fact that refrigerant will migrate through the system into the oil in the compressor. On reaching the compressor, the refrigerant is absorbed into the oil and this migration will continue until the oil is saturated with liquid refrigerant. The amount of refrigerant the oil will attract is primarily dependent on the temperature differential between the oil and refrigerant. When the pressure of a saturated mixture of refrigerant and oil is suddenly reduced, as happens in a compressor on start up, the amount of liquid refrigerant required to saturate the oil is drastically reduced, and the remainder of liquid refrigerant flashes into vapour, causing violent boiling of the refrigerant and oil mixture. This causes the typical foaming often observed in the compressor on start up, which can move all of the oil out of the compressor in less than a minute. With longer piping run, more refrigerant charge is required, thus causing more refrigerant to migrate into the oil. Foaming will be prolonged during start-up causing higher probability of the compressor failure. The introduction of excessive liquid refrigerant into the compressor can also cause a loss of oil pressure or oil delivery to the bearings even though the level of the refrigerant and oil mixture in the compressor is high. The high percentage of liquid refrigerant entering the compressor not only reduces the lubricating quality of the oil but on entering the oil pump the intake may flash into vapour, restricting the entrance of adequate oil to maintain proper lubrication of the compressor bearings. Should this oil dilution effect continue, compressor failure will occur. 1-3

10 1.3 Compressor Failur ailure With longer piping installations, the chances for compressor failure to occur becomes higher. The following are the common causes of mechanical failure due to long piping applications: a) Refrigerant flood back. In long piping application, the system will need to have a higher refrigerant charge level in order to obtain reasonable capacity. Therefore, the system will accumulate more liquid refrigerant. Refrigerant flood back occurs when liquid refrigerant flows through the suction line into the compressor during the running cycle. The liquid refrigerant will wash away the oil off the bearing surface and result in excessive wear. b) Flooded start. Liquid migration happens when the compressor is off for long periods. Refrigerant migrates to the compressor and mixes with oil. During start-ups, refrigerant foaming will wash the oil away from the bearings. With longer pipe length, more refrigerant gets to migrate into the oil and the foaming becomes more violent during start-ups. The moving parts get starved of lubrication during this period of time, and this will cause failure. c) Slugging. Slugging is the result of significant quantity of liquid refrigerant entering into the cylinder of the compressor. The hydraulic force from liquid refrigerant, oil or a mixture of both will damage the compressor cylinder. In short, slugging results from severe flooded starts where some of the foam gets sucked into the compression chamber, resulting in liquid compression. The risk of slugging is higher with long piping installations because of the higher refrigerant charge required by the system. d) Loss of oil. With long piping applications, there is a higher risk of the compressor oil flowing through the system and being trapped within the system (evaporator, condenser, piping, accumulator and other components) and returning only a little oil to the compressor. Lack of lubrication will lead to: i. Oil not reaching the bearings. ii. Oil dilution. iii. Oil thinning by overheating. A symptom of this problem is the compressor gets overheated due to the friction in moving parts. 1-4

11 1.4 High Installation tion Cost Consider the equation below: Q = AV... (1.3) where, Q = Flow rate, m 3 /s. A = Cross-sectional area of pipe (based on Inside Diameter), m 2. V = Average velocity, m/s. When a smaller diameter pipe is used, high velocity is required to convey the necessary quantity of fluid. But based on equation (1.1), high velocity causes larger pressure drop and hence, the capacity will be reduced drastically. It will also increase the operating cost due to the compressor having to do more work. When larger pipes are used, a lower velocity is required to give the desired quantity of flow. Low velocity will create less pressure drop. However, from the stand point of initial cost, the larger pipes are more costly than the smaller pipes. When applying long piping, the following should be noted by end users (which will increase the cost of installation): a) Longer pipe used. b) Bigger pipe used to reduce the pressure drop. c) More refrigerant charge. d) More installation work. e) More potential problems encountered (flood back, slugging, oil return and etc). 1-5

12 1.5 Unit Orientation tion The location of both indoor unit and outdoor unit are very important. Different types of orientation in different operating mode will have different effects on oil return, refrigerant migration, and liquid refrigerant entering the compressor. The following section will show what could happen if : a) The indoor unit is below the outdoor unit, or b) The indoor unit is above the outdoor unit. Cooling mode oil return case. Consider when the indoor unit is below the outdoor unit. When the system is operating, the oil will have to go back upwards to the compressor against gravitational forces. The amount of oil return may be very little and will directly harm the compressor internal moving parts. If both the ambient temperature for both indoor and outdoor are low (e.g. 19 C or lower), the oil viscosity will increase and this makes the amount of oil returning back to the compressor much lesser. This is a very critical condition with the outdoor unit located above the indoor unit. With long piping lengths, this phenomenon becomes more severe. See Figure 1.1. oil return up O/D [cond] 19ºC low fan I/D [evap] 19ºC Figure 1.1 : Oil return - cooling mode 1-6

13 Cooling mode cold start case For the refrigerant migration case, the vapor refrigerant will tend to move to the compressor when it is off for long periods. This phenomenon is explained in Section Note that this condition becomes more apparent when there is a larger ambient temperature difference between the indoor and outdoor units. When the indoor unit is above the outdoor unit, the migration of refrigerant becomes easier as the downward flow of refrigerant is helped by gravity. See Figure 1.2. With longer pipes, the degree of migration increases due to the higher refrigerant charge. I/D [off] 35ºC O/D refrigerant migration [off] 19ºC or lower Figure 1.2 : Cold start cooling mode Cooling mode liquid flood back case When the system is operating in cooling mode, any unvaporized liquid refrigerant will flow out from the evaporator unit and flood into the compressor. Generally, a compressor cannot tolerate any liquid flow into it as liquid compression will occur. This will damage the compressor moving parts. This phenomenon is more critical when the indoor unit is higher due to the gravitational forces making the liquid easier to flow downwards. Figure 1.3 represents this case. This situation is aggravated with long piping installations due to the extra refrigerant charge required. low fan I/D [evap] 27ºC or 19ºC O/D liquid easier to flow down to compressor [cond] 46ºC or 19ºC Figure 1.3 : Liquid flood back - cooling mode 1-7

14 Heating mode oil return case When running under the heating mode, the connecting pipes become the refrigerant discharge lines. By having the indoor unit (condenser) on a higher elevation with respect to the outdoor unit, oil will be pumped upwards by the compressor. Figure 1.4 represents this case. With long pipe installation, the oil velocity may become lower and gets accumulated inside the indoor heat exchanger. oil pumped up I/D [cond] 20ºC O/D [evap] 8ºC or lower Figure 1.4 : Oil return - heating mode Heating mode cold start case For the refrigerant migration case, the same phenomenon as explained in Section will occur. When the indoor unit is above the outdoor unit, the migration of refrigerant takes place much easier due to gravitational effects. This case is more critical than the Cooling Mode (section 1.5.2) as the outdoor unit ambient temperature is much lower, causing a higher temperature differential between the two indoor and outdoor units. Figure 1.5 represents this case. The longer the pipe is, the higher will be the migration rate due to the higher refrigerant charge. I/D [off] 20ºC refrigerant migration O/D [off] 8ºC or lower Figure 1.5 : Cold start heating mode 1-8

15 Heating mode liquid flood back case When the indoor unit (condenser) is higher than the outdoor unit (evaporator), and the system is running under the heating mode, the liquid refrigerant from the condenser is able to flow down to the outdoor unit more easily with the help of gravitational pull. With a high refrigerant charge, there is a higher risk of having liquid compression. Therefore, due to this type of unit orientation, and also because of the defrost cycle (see Section 1.5.7), it is considered as a critical condition for heating cycle. The longer the pipe is the higher will be the risks due to the higher refrigerant charge. I/D [cond] 27ºC O/D liquid easier to flow down to compressor [evap] 24ºC Figure 1.6 : Liquid flood back heating mode Defrosting cycle The defrosting cycle occurs when the system is running under the heating mode. The purpose of this cycle is to help melt any ice build-up on the outdoor coil which has been operating as an evaporator. If the ice is not removed, the heating performance of the system will deteriorate as the ice will act as an insulation on the coil surface, preventing heat transfer. To defrost, the system will momentarily switch back to the cooling mode (when the operation of the 4-way valve reverses the refrigerant flow) where the outdoor coil becomes the hot condenser to melt the ice. When the defrost is completed, the system will then resume back to heating mode. During this defrost period, the indoor unit fan will stop. As a result, this may cause the liquid refrigerant entering the indoor coil to not be able to evaporate fast enough. The excess liquid refrigerant will then flow to the compressor. By having the indoor in a higher elevation and with long pipe length (hence, higher refrigerant charge), this becomes more evident as the liquid flowing down is assisted by gravity. Liquid flood back and slugging may occur. This in turn will lead to compressor failures. See following diagrams. 1-9

16 Indoor coil (condenser) Expansion device 4-way valve Ice build-up on coil surface Outdoor coil (evaporator) Compressor Figure 1.7 : Refrigerant flow during normal heating cycle Fan stopped Indoor coil (evaporator) Incomplete evaporation, liquid floodback to compressor Expansion device 4-way valve Compressor Outdoor coil (condenser) Figure 1.8 : Refrigerant flow during DEFROST cycle 1-10

17 1.6 Piping Length Limits In summary, compressor failure is the main consideration with long piping applications. See Appendix A for more detailed explanation on compressor failures. Piping length should be as short as possible to prevent the compressors from breaking down. In view of the importance of unit orientation, care must be taken to keep to the allowable maximum height difference. Special precautions must be taken when there is no choice but to use longer pipe lengths. Understandably, the extent of these precautions must be balanced with the increased cost of installation. The following table shows the recommended piping length limits for different condensing units: Model Maximum Maximum Maximum number of Length (m) Height (m) bends allowed SL10B/BR/10C/CR SL15B/BR/15C/CR SL20B/BR/20C/CR* SL25B/BR/25C/CR* SL30/40/50C/CR* SL60/61C/CR* VCU25A VCU30/35/40/50A MSS60C/CR/75C/CR* MSS100/125/150B/BR/C/CR/D/DR/E/ER* MSS200/250/300D2/DR2* Note* : Applicable to model with built-in accumulator for long piping application only. Please refer to distributor for further details. Table 1.1: Pipe lengths for different condensing units The standard pipe length, where the units are rated at is 7.6 meters (25 feet). A more thorough explanation on precautions needed with longer pipe installations are given in the next chapter. Pipe lengths longer than those specified in Table 1.1 is not recommended. It is also not necessary to change the recommended pipe sizes for these units, as long as these limits are not exceeded. In general, the outdoor unit pipe size connections should always be used as the reference. Table 1.2 shows the pipe sizes for different condensing units. 1-11

18 Model Pipe size ** Suction ( ) Liquid ( ) SL10B/BR/C/CR 3/8 1/4 SL15B/BR/C/CR 1/2 1/4 SL20B/BR/C/CR 5/8 1/4 SL25B/BR/C/CR 5/8 3/8 VCU25/30A 5/8 3/8 SL30C/CR 5/8 3/8 SL40/50C/CR 3/4 3/8 VCU35/40/50A 3/4 3/8 SL60/61C/CR 3/4 1/2 VCU60A 7/8 1/2 MSS60C/CR 3/4 1/2 MSS75D/DR 1 1/2 MSS100B/BR/C/CR/D/DR/E/ER 1-1/8 5/8 MSS125/150B/BR/C/CR/D/DR/E/ER 1-3/8 5/8 MSS200D2/DR2 1-1/8 5/8 MSS250/300D2/DR2 1-3/8 5/8 Note ** : Subject to change without prior notice. For updated information, please refer to Technical Manuals. Table 1.2 : Pipe sizes for different condensing units 1-12

19 Section 2 Precautions On Long Piping Installations There are several considerations when dealing with long piping installations. These include: a) Additional oil. b) Additional refrigerant. c) Oil traps. d) Suction accumulator. e) Oil separator. f) Pipe sizing. g) Crankcase heater. h) Pump down cycle. i) Minimise bends. j) Minimise excessive heights. 2-1

20 2.1 Additional ditional Oil Each compressor has a rated oil holding capacity. It is important to check with the compressor specifications on these oil charges as each compressor comes pre-charged with oil. This specified compressor oil holding charge is sufficient for a standard piping installation of 7.6m (24.9 feet). With longer pipe lengths, it is important to add in additional oil into the system. This is necessary because some of the oil will be pumped out of the compressor and stick to the internal pipe surfaces. Additional oil is needed to maintain a safe oil level in the compressor sump. As a guideline, from the actual pipe run, for every 10ft of extra length from standard testing length, 3 fl.oz of oil (1 fl.oz 30cm l) should be added into the system. Sample calculation 01: a) Standard factory testing length = 7.6m / 24.9 ft b) Actual pipe length = 10.0m / 32.8ft c) Compressor specifications: Oil charge = 38fl.oz. Extra length = 32.8ft 24.9ft = 7.9 ft With 10ft = 3fl.oz., the extra oil charge for the extra length of 7.9ft = 2.37fl.oz. Therefore, an additional 2.37fl.oz of oil should be added into the compressor in this example, giving a total charge of = fl. oz. 2-2

21 2.2 Additional ditional Refrig efriger erant Similar to the additional oil case, each compressor has a rated refrigerant maximum holding capacity. As a general practice, do not exceed these limits in order to prevent the compressor from breakdown. It is important to double check with the compressor specifications on the refrigerant charge limits. With the longer pipe lengths, more refrigerant is needed to fill the extra volume. The refrigerant charge in the delivered unit is sufficient for the standard pipe length of 7.6 meters. It is also important to understand the difference between having the expansion device located in the indoor unit or the outdoor unit. If the expansion device is located at the indoor unit, the entire liquid line will be filled with liquid refrigerant before expanding at the indoor unit. This will be considered as a single phase flow in the liquid line. The refrigerant amount required will be significantly higher. When the expansion device is in the outdoor unit, the refrigerant will begin to expand along the liquid line. This will be considered as a two phase flow in the liquid line. As a result, the refrigerant amount required is lower. The additional amount of refrigerant required can be calculated once the density of the refrigerant along these pipes is known. As a guideline, this value should not be more than the maximum refrigerant charge of the compressor. Sample calculation 02: a) 1 st option pipe length = 40ft = 12.2m b) 2 nd option pipe length = 75ft = 22.9m c) Standard pipe length = 7.6m d) L type hard copper tube with OD size = ½ e) Density of refrigerant R-22, ρ R22 at 40 o C = 1129kg/m 3 (saturated temperature of liquid at condenser outlet). f) Standard factory charge = 2.50kg g) Compressor maximum refrigerant charge = 3.60kg h) Given that total charge must not be greater than the compressor maximum charge. i) The expansion device located at the indoor unit. From Table D.5, we look at L type hard copper tube with OD = ½, ID = 0.430in ( 0.011m). Also from the R22 Thermo Physical Properties, ρ R22 at 40 o C = 1129kg/m 3. It is known that: Mass = Volume x Density Hence, m = V x ρ R22 (2.1) Where, m is the additional refrigerant mass required for the additional pipe length V is the additional refrigerant volume due to the additional pipe length If d is the pipe internal diameter L is the pipe length which is equal to the on site pipe length minus the standard pipe length Then, m = 0.25 x π x d 2 x L x ρ R22 m L=12.2m = 0.49kg (for 1 st option pipe length) m L=22.9m = 1.64kg (for 2 nd option pipe length) For 1 st option case; Total charge = Extra charge + Standard factory charge = = 2.99kg! 2.99kg < 3.6kg For 2 nd option case; Total charge = Extra charge + Standard factory charge = = 4.14kg! 4.14kg > 3.6kg Thus, the second option pipe length is not recommended. 2-3

22 Sample calculation 03: Continuing from sample calculation 02, but this time with an outdoor expansion device. As explained earlier this will be considered as a two phase flow in the liquid line. The refrigerant amount required would now be lower. It is difficult to exactly quantify the refrigerant charge in this situation as the actual expansion process that occurs in the expansion device is not known. However, it is estimated that the refrigerant required is 40% less compared to the amount needed when the expansion device is at the indoor unit. Hence, m L=12.2m = 0.49/1.4 = 0.35kg (for 1 st option pipe length) m L=22.9m = 1.64/1.4 = 1.17kg (for 2 nd option pipe length) For 1 st option, the total charge = Extra charge + Standard factory charge = = 2.85kg ( < 3.6kg ) For 2 nd option, the total charge = Extra charge + Standard factory charge = = 3.67kg ( > 3.6kg ) The second option pipe length is still not recommended in this situation. 2-4

23 2.3 Oil Trap In normal installation cases, oil traps are not required. However, when piping is long, oil traps are required to be installed at fixed intervals along the vertical suction pipe. This is especially so when the outdoor unit is located on a higher elevation than the indoor unit. These oil traps help to get any accumulated oil to move upwards, as shown in Figure 2.1. The design of the traps will forcibly cause the gaseous refrigerant to pass through the oil thus, carrying it upwards back to the compressor. As a guideline, an oil trap is required at every 10 to 15ft (3 4.6m) intervals. Figure 2.2 illustrates how oil traps are installed. However, the high pressure drop across such traps may cause high capacity reduction. Figure 2.1: Internal refrigerant flow inside an oil trap TO COMPRESSOR TO COMPRESSOR EVAPORATOR 15 EVAPORATOR Figure 2.2: Oil traps installations 2-5

24 2.4 Suction Accumula ulator A suction accumulator serves as a vessel to store any liquid refrigerant which may return back along the suction line to the compressor. It will protect the compressor from liquid floodback and slugging. All OYL heat pump units have a built-in suction accumulator in the outdoor unit, but not for the cooling only units. Nevertheless, it is not recommended to have an additional suction accumulator installed in the units as long as the maximum pipe length limits defined in Section 1.6 are not exceeded. The existing accumulator in the unit is able to provide sufficient protection to the compressor within the specified pipe length limits. A more thorough explanation and study on suction accumulator is given in Appendix B. 2.5 Oil Separ parator The oil separator has been designed to remove compressor oil along the discharge line. It is useful to ensure sufficient oil return back to the compressor. However, it is not recommended that an additional oil separator be installed in the units as long as the maximum pipe length limits are not exceeded. The unit will be able to function properly within the specified limits, provided that care is taken to have oil traps along the pipe line (see Section 2.3) and that the refrigerant velocity in the pipes are sufficiently high to carry the oil back to the compressor (see Section 2.6). Appendix C gives a more detailed study and explanation of the oil separator. 2-6

25 2.6 Pipe Sizing Refrigerant pipe lines must be selected for optimum size with respect to: a) Initial cost. b) Pressure drop. c) Oil return. It is desirable to have line sizes as small as possible from the standpoint of low initial cost. However, the overall system performance must be evaluated and the following recognised: a) Suction and discharge line pressure drop due to friction losses reduces compressor capacity and increases power consumption. b) Liquid line pressure drop due to friction loss and liquid static head may cause flash gas (Flash gas is the refrigerant gas which results from the vaporisation of liquid refrigerant to cool the remaining liquid refrigerant to a lower pressure level). c) Suction and discharge lines must be sized for proper gas velocity to assure oil return to the compressor. The design considerations with long refrigerant piping installations are: a) Assure positive and continuous return of oil to the compressor crankcase. b) Refrigerant pressure losses are inevitable with long piping. This should not be remedied at the expense of retarding oil return to the compressor. c) Prevent liquid refrigerant from entering the compressor during running, off cycles and start up. d) Avoid trapping of oil in the evaporator or suction line which may subsequently return to the compressor as large slug with possible damage to the compressor. In view of the above mentioned considerations, it is recommended that the refrigerant pipe size be maintained as specified on the indoor and outdoor units. It is not necessary to resize the pipe as long as the refrigerant pipe length limits are not exceeded. 2-7

26 2.6.1 Suction line The suction line should have the following characteristics: a) A total pressure drop of not more than 2 o F change in saturation temperature (which is equivalent to 3 psi for R-22 40ºF evaporating temperature). Of course, for long piping installations, this value will be much higher. b) Sufficient velocity (>1500fpm for vertical riser, >750 fpm for horizontal runs) for oil return to the compressor (greater refrigerant velocities are obtained by decreasing the size of the suction line. However this will create a higher pressure drop). Note: (Criteria (b) has higher priority over (a)). c) Prevention of liquid refrigerant from draining into the compressor during OFF cycle. When installing the evaporator below the compressor, using a trap at the bottom of the pipe riser is necessary. The purpose of the trap is to drain oil and liquid refrigerant out of the line to which the expansion valve bulb is strapped. See following diagram: Trap To compressor Evaporator Figure 2.3: Evaporator located below the compressor. When installing the evaporator above the compressor, using an inverted Loop can prevent refrigerant from draining into the compressor during OFF cycle. However, the loop will not prevent refrigerant migration due to temperature of the evaporator being higher than the compressor. Evaporator To compressor Figure 2.4: Evaporator located above the compressor. The trap should be as small as possible to prevent large slugs of oil being returned to the compressor when the trap clears. When a suction riser is 30 feet or more in length, an oil trap should be installed every 15 feet of vertical rise. This trap aids in oil return and provides a drainage point for oil, which is en route up the riser when the compressor stops. When the unit starts again, the oil is returned to the compressor quickly and in a relatively small slugs. See Section

27 2.6.2 Discharge line Pressure drop due to friction loss in discharge lines causes the compressor to operate in higher pressure resulting in reduced capacity and increased power consumption. Discharge lines should have: a) A total pressure drop of 1 o F to 2 o F change in saturation temperature (equivalent to 3.5 psi for R ºF condensing temperature). Long piping will give higher readings. b) Sufficient velocity (>1500fpm) for oil return to the compressor. Note: (Criteria (b) has higher priority over criteria (a)) c) A means to prevent refrigerant from draining back to compressor head during the OFF cycle. Long discharge pipe lines are only encountered during the reversed cycle heating mode. The same pipe is the suction line during the cooling mode. Therefore, the main consideration with long discharge pipe installations is the discharge oil line trap, which is actually the same as the suction line oil traps. The same rule applies, i.e. a trap should be installed every 15 feet of vertical rise. These traps will aid in oil return and provide a drainage point for oil that is en route up the riser when the compressor stops, as well as for liquid refrigerant which may condense during the OFF cycle Liquid line When the refrigerant is in liquid state, the oil in the liquid line is readily carried along by the refrigerant to the evaporator. There is no problem with oil return in liquid lines. Thus, the design of the liquid piping is less critical than that of the suction lines and the discharge lines. The problem encountered in the liquid line is mainly one of preventing the liquid from flashing before it reaches the refrigerant control (capillary tube or thermal expansion valve). The problem of flash gas in the liquid line is that: a) It reduces the capacity of the refrigerant control. b) It causes erosion of the valve pin and seat. c) It often results in erratic control of the liquid refrigerant to the evaporator. To avoid flashing of the liquid in the liquid line, sufficient liquid subcooling is required along the pipe line. The liquid line should be designed with a maximum gas velocity of 360 fpm. 2-9

28 2.6.4 Refrigerant piping checking method This section provides a general guide on how to check the refrigerant piping. Procedure: 1) Select indoor and outdoor models: a) Capacity (Btu/hr or tons) b) Original gas and liquid line size. c) Standard factory length. 2) Obtain the following parameters: a) Horizontal pipe length. Check if pipe limits have been exceeded. b) Vertical pipe length. Check if pipe limits have been exceeded. c) Number of bends. d) List of fittings installed, e.g.: filter drier, valves, sight glass. 3) Based on the information from 1(a) and 1(b), use the pressure drop charts (Figure D.4) and velocity charts (Figure D.3) for the following: a) Pressure drop for current pipe size and pipe size at least 2 sizes bigger than current size. b) Velocity in pipe for current pipe size and for pipe size at least 2 sizes bigger than current size. Note: The charts are applicable only for refrigerant R-22. 4) Based on the information from 2(a) and 2(b), calculate the actual pressure loss for the pipe sizes. 5) Based on the information from 2(c) and 2(d), find from Table D.1 to Table D.6 for the equivalent lengths of the fittings such as elbows and valves. 6) Check the values against the design criteria set for pressure drop and velocity in gas line. The design criteria are listed as below: a) Gas line criteria: i) Minimum horizontal line gas velocity = 750fpm ii)minimum vertical risers gas velocity = 1500fpm 7) Check the refrigerant velocity of the standard pipe size. Check also the refrigerant pipe pressure drop. Determine the performance of the standard pipe size. Compare the values with those pipes which are 2 sizes larger as reference. 8) Get factory standard length; calculate the additional oil from the actual pipe run. For every 10ft of extra length from standard testing length, add 3 fl.oz. of oil (1 fl.oz. 30cc). 9) Refer to Table D.5. Based on the length and inner diameter of the pipe, find the amount of refrigerant to be added in. 2-10

29 Sample calculation 06: Capacity = 50000Btu/hr (Note: 1ton = 12000Btu/hr). Suction line size = ¾ (OD). L type hard copper pipe. Liquid line size = 3/8" (OD). L type hard copper pipe. 40 o F (4.4ºC) evaporation temperature. 120 o F (48.9ºC) condensing temperature. (Refrigerant density = 1087 kg/m 3 ) Standard testing length = 7.6m. Compressor maximum R22 holding capacity = 4.5kg. Standard R22 charge = 2.5kg. System is utilizing expansion valve (indoor). Cooling mode system. 6 Evaporator Condenser 4 NOTE : Dimensions are not to scale Figure 2.5 : Schematic diagram for sample problem Capacity = 4.17 tons (5 hp) Total pipe length = 68 ft (20.7 meters) Elevation = 15 ft (4.6 meters) With reference to the specifications in Secton 1 for a 5 hp unit (50 meters maximum length, 15 meters maximum height), the total pipe length and elevation are still within the limits. 2-11

30 We can thus build the following tabulation: Section OD Size (inch) Velocity (fpm). Refer to Figure 2.6. Number of bends, a. Equivalent length for 1 bend, b (Note: Assume long radius 90 o bend). Refer to Table D.1 & D.2. Equivalent length for all bends c = a x b Misc. fittings equivalent length (ft), d. Refer to Table D.3, D.4 & D.6. Total length (ft), e. Total equivalent length (ft) f = c + d + e Pressure drop / 100ft (psi), g. Refer to Figure 2.7 Calculated pressure drop (psi) h = f x (g/100) Suction Liquid 3/ / / / / Table 2.1: Tabulated values for various parameters. For this calculation example, it is necessary to perform an interpolation to obtain the equivalent length l e (7/8) and l e (5/8) of the bends from the available data in Table D.1 and D.2. Formula used to obtain these values is: l i = l i-1 + [(φ i φ i-1 ) /(φ i+1 φ i-1 )] x (l i+1 l i-1 ) Where: l i = Equivalent length to be determine. l i-1 = 1 step smaller equivalent length. l i+1 = 1 step bigger equivalent length. φ i = Diameter of l i. φ i-1 = 1 step smaller diameter. φ i+1 = 1 step bigger diameter. Equivalent length for 90 o long radius bend for 7/8", l e l e (7/8) = l e (3/4) + [(OD 7/8 OD 3/4 )/(OD 1 OD 3/4 )] x (l e (1) l e(3/4) ) = [(7/8 3/4)/(1 3/4)] x ( ) = 1.55 Equivalent length for 90 o long radius bend for 5/8", l e l e (5/8) = l e (1/2) + [(OD 5/8 OD 1/2 )/(OD 3/4 OD 1/2 )] x (l e (3/4) l e(1/2) ) = [(5/8 1/2)/(3/4 1/2)] x ( ) = 1.20 With these answers, the pressure drop across the pipes can be calculated, as shown in Table 2.1. As expected, the pressure drop across the larger pipes is found to be much lower. With this installation, the extra pipe length Therefore, the extra lubricant oil required = 68ft (7.6m / (0.3048m/ft)) = 43.1ft = Extra length x (0.3fl.oz/ft) = 12.92fl.oz. 2-12

31 Refer to Table D.5 to determine the internal pipe diameter. The extra refrigerant required is calculated as follows : OD =3/8" case (ID = 0.305"). Extra R22 = ρv = 0.25πd 2 Lρ = 0.25(3.142)(0.305in x (0.0254mm/in)) 2 (68ft x (0.3048m/ft) 7.6m)(1087kg/m 3 ) = 0.67kg Total charge = = 3.17kg (OK) OD = ½ case (ID = 0.43"). Extra R22 = ρv = 0.25πd 2 Lρ = 0.25(3.142)(0.430in x (0.0254m/in)) 2 (68ft x (0.3048m/ft) 7.6m)(1087kg/m 3 ) = 1.34kg Total charge = = 3.84kg (OK) OD = 5/8" case (ID = 0.545"). Extra R22 = ρv = 0.25πd 2 Lρ = 0.25(3.142)(0.545in x (0.0254mm/in)) 2 (68ft x (0.3048m/ft) 7.6m)(1087kg/m 3 ) = 2.15kg Total charge = = 4.65kg (NG) Because the pipe length limit is not exceeded, the original pipe size is still recommended to be used in this system. The calculations reveal that the system will be safe to operate, but additional oil of 12.9 fl. oz. and additional refrigerant R22 charge of 0.7kg must be added. In this example, R-22 is used as the refrigerant. When other refrigerants are used, the corresponding refrigerant properties must be applied. For example, properties of refrigerant R407C and R410A are available in the Appendix section. The corresponding pipe charts must also be used in the computation. 2-13

32 Figure 2.6: R-22 refrigerant velocity chart 2-14

33 Figure 2.7: R-22 refrigerant pressure drop chart 2-15

34 2.7 Crankcase Heater Crankcase heater is a sealed heater installed with close contact to the outer circumference at the bottom of the compressor. Examples of crankcase heater and how crankcase heater is installed onto a compressor is shown in Figure 2.8, Figure 2.9, and Figure Figure 2.18: Example of crankcase heater. Figure 2.19: Crankcase heater installed onto a rotary compressor. Figure 2.10: Crankcase heater installed onto a reciprocating compressor. 2-16

35 The purpose of installing crankcase heater is to protect the compressor from the negative effects of the liquid refrigerant in the compressor and in the lubricating oil. In simple words, crankcase heaters are frequently used to retard migration. It removes the refrigerant by heating from the outside. Refrigerant entering the compressor will be vaporized and driven back into the suction line. Crankcase heater should be installed during long piping installations where the risk of liquid refrigerant migration is much higher. The crankcase heater size differs according to the capacity and the application of the compressor. A heater of about 40W to 80W should be used. When the compressor is inactive for a long period, the crankcase heater should be energized for at least 6 to 12 hours before operation of the compressor is started. Please note that burning might occur if oil gets on to the crankcase heater. Deterioration due to water condensation and acoustic insulation materials (such as pheuol products ) can lead to defective insulation. 2.8 Pump Down Cycle le The most positive dependable means of properly controlling the liquid refrigerant, particularly if the charge is large, is by means of a pump down cycle. By closing a liquid line valve, the refrigerant can be pumped into the condenser and receiver, and the compressor operation controlled by means of a low-pressure control. The refrigerant can thus, be isolated during periods when the compressor is not in operation. Migration to the compressor and crankcase is prevented. Although the pump down cycle is one of the protection method against migration, it will not protect against liquid flooding during operation. 2-17

36 2.9 Minimize e Bends Piping between the condenser and evaporator units shall not have too many bends. Bends should be avoided as much as possible. When the number of bends (bending angle) is large, the internal pipe resistance increases, and the refrigerant flow is impaired. These bends tend to retard oil return. The compressor capacity is also reduced and there are higher risks of compressor failures. Refer Section 1.6 for the recommended maximum number of bends. Figure 2.11: Too many bends 2-18

37 2.10 Minimize e Excessi cessive e Heights The system does not perform correctly when both the condenser and evaporator units are too far away from each other (either vertically or horizontally). The required refrigerant quantity increases and the products guaranteed range is exceeded. Also, the circulation of refrigerant and lubrication oil malfunctions, the capacity drops, and compressor trouble may occur. The piping length should be as short as possible because the capacity and the reliability decreases as piping length increases. Select the shortest length possible. Refer Section 1.6 for the recomended maximum heights and lengths. Figure 2.12: Too long horizontal length. Figure 2.13: Excessive height different. 2-19

38 Section 3 Capacity Effect One of the most frequently asked question with long piping installations is How much is the effect to the cooling or heating capacities when operating with long piping? It has been mentioned that with longer piping, the capacity will be lower due to pressure losses along the pipe lines. However, to determine the magnitude of the capacity reduction is not just a simple matter of calculating the pressure drops along the pipe line. The cooling or heating capacity of a system is very much dependent on the operating suction and discharge pressures of the compressor. It is also dependent on the amount of superheat at the suction, and subcool at the condenser outlet. All of these boil down to the refrigerant mass flow rate pumped by the compressor. When the pipe lines become longer, and the refrigerant charge is increased, the values of both the superheat and subcool will also change. This will affect the mass flow rate. In other words, in order to accurately determine the capacity, we will need to measure the pressures, superheat, and subcooling. Another consideration is that with the longer pipe lines, there will be higher heat losses due to conduction along the pipe length. Of course, this can be overcome by ensuring good pipe insulation on the pipes. Generally, a suitable pipe insulation of sufficient thickness (e.g. Superlon/Armaflex, ½ thick) will be effective in giving good thermal insulation. In view of the above, it is difficult to determine accurately the capacity reductions due to long pipe installation. It is not practical to directly measure both the compressor suction and discharge pressures as there are no pressure taps at these locations in the air-conditioning unit. A table or graph of the compressor performance is also required to determine this value. Nevertheless, it is possible to roughly estimate the capacity by using the calculated pressure drop along the suction and discharge pipe lines in relation to the original rated capacity of the system with the standard pipe length. It is assumed that the rated evaporating and condensing temperatures of the system remain Note 1. the same. Note 1: Strictly speaking, this is not true, as a reduction in suction pressure will also reduce the discharge pressure (and viceversa). The system balancing will be affected. However, this assumption has been made to simplify the estimation process. 3-1

39 Cooling Capacity The cooling capacity of the unit has been rated with a standard pipe length at a specific evaporating temperature and condensing temperature. Suction pipe Ps P Suction pressure Evap. temp. (sat.), Te Cond. temp. (sat.), Tc P Liquid pipe (2- phase) )))) )))) Std. pipe length Figure 3.1: Cooling cycle With a longer pipe length, the compressor suction pressure must become lower in order to maintain the same evaporating temperature at the coil, i.e. Te. As a result, the refrigerant mass flow rate of the compressor reduces, giving a lower cooling capacity. The longer the pipe the lower the suction pressure. The pressure drop along the pipe can be expressed as equivalent temperature reading because of the general acceptance of this method of pipe sizing. The corresponding pressure drop in psi (or kpa) may be determined by referring to the saturated refrigerant properties. Different refrigerants will give different values of pressure drop. The following general capacity trend has been extracted from CARRIER Handbook: Suction pipe line Pressure Drop Compressor Capacity, % No Line Loss 100 2ºF (1.1ºC) Line Loss ºF (2.2ºC) Line Loss 92.2 Table 3.1: Suction line pressure drop against compressor capacity 3-2

40 Heating Capacity With the action of the 4-way reversing valve, the suction pipe will now become the discharge pipe. Because of the expansion device configuration, the liquid pipe will maintain the same 2-phase flow. Discharge pipe Discharge pressure P Pd Cond. temp. (sat.), Tc Evap. temp. (sat.), Te P )))) Liquid pipe (2-phase) )))) Outdoor unit Std. pipe length Indoor unit Figure 3.2: Heating cycle As before, the heating capacity has been rated at specific evaporating and condensing temperatures, with the standard pipe length. With a longer pipe installation, there will be no change in the suction pipe line pressure drop, but rather the effect comes from the discharge pipe. Due to the additional discharge pipe pressure drop, the compressor discharge pressure has to be increased in order to maintain the same condensing temperature at the coil. As a result of this higher discharge pressure, the compressor capacity will decrease, increasing the power input. The following table illustrates this situation: Discharge pipe line Pressure Drop Compressor Capacity, % No Line Loss 100 2ºF (1.1ºC) Line Loss ºF (2.2ºC) Line Loss 96.8 Table 3.2: Discharge line pressure drop against compressor capacity. Notice that the amount of capacity loss due to the discharge pipe line pressure drop is lower than the equivalent pressure drop along the suction line. 3-3

41 3.1 Pressur essure e Drop op Charts ts The following explains the refrigerant pipe line pressure drop chart which has been described under the topic Pipe sizing in Section 2. A C B Figure 3.3: Pressure drop chart for R

42 Figure 3.3 shows the pressure drop chart for R-22 This chart gives the pressure drop along 3 different types of refrigerant pipe lines, i.e. a) suction pipe [A] b) liquid pipe [B] c) discharge pipe [C] The differentiations can be found on the right hand side of the chart. Note that for each of the liquid and discharge pipe, there is only one representative line irrespective of the pipe size. But there are several lines for the suction pipe, corresponding to different evaporating temperatures. Figure 3.3 is applicable only for R-22 refrigerant. Other types of refrigerant will require different charts. The chart for R-407C refrigerant has also been included in the Appendix section. The following explains how the pressure drop chart is used. Refer to Figure 3.3: 1. Determine the cooling capacity of the system in Refrigerant Tons. 2. Project downwards from the capacity scale at the top. Intersect the three types of pipe lines [A], [B], and [C]. 3. Determine the evaporating temperature and condensing temperature of the system. 4. Project horizontally to the left at the point of intersection with the suction line (at the corresponding evaporating temperature), discharge line, and liquid line. 5. Intersect the projected lines with the left section of the chart at the corresponding pipe sizes. 6. Then, project downwards to the pressure drop scale at the bottom of the chart. 7. Follow the slope lines to read from the bottom scale for the desired condensing temperature. Note that the pressure drop values are given in the unit psi/100 Feet. 3-5

43 The next step is to estimate the capacity changes. The approach used in this manual is by referring to the capacity changes on the actual compressor performance curves. The focus will be on three different types of compressors: a) Matsushita rotary compressor b) Bristol reciprocating compressor c) Copeland scroll compressor Table 3.3 shows the average percentage of capacity loss per degree of pressure drop along the suction line for each of these compressors. This is done for cooling mode at a specific evaporating temperature range of 40 45ºF ( ºC) and condensing temperature range of ºF ( ºC). Matsushita rotary % capacity loss per ºF (R-22) % capacity loss per ºF (R-407C) 1 h.p. -2.1% -2.1% 1.5 h.p. -1.9% -2.0% 2 h.p. -1.8% -1.9% 2.5 h.p. -1.3% -1.9% Bristol reciprocating % capacity loss per ºF (R-22) 3 h.p. -2.0% 4 h.p. -2.3% 5 h.p. -2.2% 6 h.p. -1.9% Copeland scroll % capacity loss per ºF (R-22) % capacity loss per ºF (R-407C) 3 h.p. -1.9% -2.0% 4 h.p. -2.1% -2.2% 5 h.p. -1.9% -2.1% 6 h.p. -1.9% -2.0% 7.5 h.p. -1.9% -2.0% 10 h.p. -1.8% -2.0% 12.5 h.p. -1.8% -1.9% Table 3.3: Average percentage capacity loss per degree pressure drop in cooling mode 3-6

44 Table 3.4 shows the average capacity losses during heating mode for every degree of pressure drop along the hot gas discharge pipe line. This is applicable for evaporating temperature range of 26 30ºF (-3-1ºC) and condensing temperature range of ºF ( ºC). Matsushita rotary % capacity loss per ºF (R-22) % capacity loss per ºF (R-407C) 1 h.p. -0.7% -0.8% 1.5 h.p. -0.6% -0.6% 2 h.p. -0.6% -0.7% 2.5 h.p. -1.1% -0.7% Bristol reciprocating % capacity loss per ºF (R-22) 3 h.p. -0.7% 4 h.p. -0.8% 5 h.p. -0.8% 6 h.p. -0.7% Copeland scroll % capacity loss per ºF (R-22) % capacity loss per ºF (R-407C) 3 h.p. -0.6% -0.7% 4 h.p. -0.7% -0.8% 5 h.p. -0.6% -0.7% 6 h.p. -0.6% -0.7% 7.5 h.p. -0.6% -0.7% 10 h.p. -0.5% -0.7% 12.5 h.p. -0.6% -0.7% Table 3.4: Average percentage capacity loss per degree pressure drop in heating mode 3-7

45 By knowing the additional pressure drop along the long pipe line, the change of cooling or heating capacity can be estimated. Referring to Sample Calculation 06, from Figure 2.7 it was determined for the 5 hp unit, that the pressure drop for the ¾ suction pipe is 15 ft. per 100 ft., and for the 3/8" liquid pipe is 25 ft. per 100 ft. The unit has been rated with a standard pipe length of 25 ft (7.6 m). Thus, if the unit is installed with a total pipe length (inclusive of bends) of 68 ft, the pressure drop along the additional pipe length can be calculated as follows: Additional suction pressure drop = (68 25) ft *15/100 = 6.45 psi. Additional liquid pressure drop = (68 25) ft * 25/100 = psi. Within the scope of this manual, the effect of the liquid line pressure drop is assumed negligible. a) For refrigerant R-22, every ºF of suction pressure drop is equivalent to about 1.5 psi. Refer to R- 22 saturated tables at 40 45ºF. Hence, the additional suction pressure drop of 6.45 psi is equivalent to 6.45/1.5 = 4.3ºF. If the unit is using a Copeland scroll compressor, the capacity reduction will be about 4.3 * 1.9 = 8.2%, i.e. giving a rated capacity of * 50,000 = 45,900 Btu/hr. If a Bristol reciprocating compressor is used instead, the capacity reduction will be about 4.3 * 2.2 = 9.5%. b) Recalculating the above example by using refrigerant R-407C, every ºF of suction pressure drop is also equivalent to about 1.5 psi. Note that R-407C is an azeotropic refrigerant, and the refrigerant dew saturated temperature is used in the calculation. Refer Tables D.30 to D.33 for the refrigerant properties. The calculation of the suction line pressure drop is done using the R-407C Pressure Drop chart as in Figure 3.4. From the chart, the pressure drop of the ¾ suction pipe is 13 psi per 100 feet. Therefore, the total equivalent pipe length of 68 ft. will now give a total pressure drop of (68 25) ft * 13/100 = 5.59 psi. This is equivalent to 5.59/1.5 = 3.7ºF. With a Copeland scroll compressor (R-407C), the capacity reduction will be 3.7 * 2.1 = 7.8%, i.e. giving a rated capacity of * 50,000 = 46,100 Btu/hr. The same can be done with other refrigerants using the saturated properties of the refrigerant. 3-8

46 10 Figure 3.4 : Example on using pressure drop chart fro R

47 During heating mode, both the pipes will now be under high pressure, with the suction pipe becoming the discharge pipe. The same example will be used to work out the pressure drop along this discharge line. This example assumes the capillary tube is found in the outdoor unit with a condensing temperature of 100ºF. Again, the effect of the liquid line pressure drop is assumed negligible. Looking at the pressure drop chart, we find that the pressure drop along this is about 10 psi /100 ft. Refer to Figure 3.4. This translates to an additional discharge line pressure drop of (68 25)ft * 10 /100 = 4.3 psi a) For R-22 refrigerant at ºF, every ºF of discharge pipe pressure drop is equivalent to 3.3psi. Therefore, the additional pressure drop of 4.3 psi is equivalent to 4.3/3.3 = 1.3ºF. With a Copeland scroll compressor, this will give a capacity reduction of 1.3 * 0.6% = 0.78%. If the heating capacity of the unit is rated at 53,000 Btu/hr, the unit will thus, have a rating of * 53,000 = 52,586 Btu/hr. b) For R-407C refrigerant, every ºF of discharge pipe pressure drop is equivalent to 3.5 psi. From Figure 3.5, a pressure drop of about 7 psi per 100 feet is obtained when this example is recalculated using R-407C. Thus, the additional discharge line pressure drop is (68 25) ft * 7/100 = 3 psi. Hence, the long piping will give an equivalent pressure drop of 3/3.5 = 0.85ºF. The capacity reduction is then 0.85 * 0.8% = 0.68% for a Bristol compressor. 3-10

48 7 13 Figure 3.5: Pressure drop chart for R-407C 3-11

49 Section 4 Piping Installation The following section will provide some guidelines for refrigerant copper pipe installation, especially in relation to long piping jobs. Since the copper pipe is a flexible material, care must be taken to ensure proper installation. 4.1 Pipe Material It is recommended that Type L or Type M hard copper pipes be used to install the split type air-conditioning units. See Table D.5 for the physical properties of the pipe. Alternatively, refrigeration tubing with thinner wall thickness, may be used. The wall thickness must be sufficient to withstand a burst pressure of at least 1700 psig (11730 kpa) when used with R-22 and R-407C refrigerant. However, working with R-410A refrigerant will require a stronger pipe material to withstand the higher working pressure. Burst strength of at least 2400 psig will be required. In view of this, do not use the softer refrigeration tubing and Type M pipes. It is recommended that Type L be used for R-410A. Joining of two pipes can be done easily by brazing with a copper filler rod. For better quality joint, a filler rod with 2% silver may be used. It may also be necessary to braze the copper pipe to a brass or steel fitting. In such instances, brazing with 34% silver filler rods must be used (together with brazing flux). 4.2 Pipe Insulation It is only necessary to insulate the cold suction pipe. Do not insulate the hot liquid pipe. If the expansion device is located in the outdoor unit, the liquid pipe will have a 2-phase flow inside which is cold. This must also be insulated. Generally, this is for the smaller units (1 to 2.5 hp). However, for heat pump units, it is important for both pipes to be insulated. This is because of the cold ambient temperatures when the unit is running in heat mode. The insulation will prevent heat loss to the ambient along the hot pipe line. Insulation can be done easily by inserting the copper pipes into elastomeric insulation pipes. Examples are Armaflex and Superlon. Use the correct insulation sizes to the corresponding copper pipe size. Do not use a larger insulation as this will create an air space which will then create condensation (sweating). Cut sections of the insulation must be glued or taped together over the pipes, e.g. at bends and joints. Recommended insulation: k-value of W/m.K Minimum insulation thickness: ½ (12.7 mm) Do not insert two copper pipes into a single large pipe insulation. Such practice will cause the system to lose performance due to heat gain or heat loss because the pipe surfaces are not in good contact with the insulation. Cross heat transfer between a cold and hot pipe can occur due to the close proximity of the two pipes. Potential sweating problems may also occur due to the created air space within. 4-1

50 4.3 Horizontal ontal Runs Normally, refrigerant pipes are run above the ceiling space. In order to do such horizontal runs, it is necessary to have supports at certain intervals so that the pipes do not sag. Supports in the form of saddles or angle brackets may be used. Multiple pipes can share the same support. ceiling hanger rods embedded into ceiling with wall plugs SADDLE SUPPORT hanger rods (threaded) copper pipes saddle bar nuts Figure 4.1: Pipe support using saddle copper pipes ANGLE BRACKET wall bracket Figure 4.2: Pipe support using angle bracket 4-2

51 It is not recommended to run the pipes on the floor, for the simple reason that people may just step on the pipes and damage them. However, should there be a need to do so; some kind of protection must be given. An example is to place the pipes into a GI trunking box which is mounted (screwed) onto the floor. Trunking box Floor level Figure 4.3: Pipe support using trunking box The following table gives recommendation for the support spacing of the copper pipes: Pipe diameter, OD ( ) Distance between spacing, (ft) Up to 5/8 6 7/8 to 1-1/ /8 to 2-1/8 10 Table 4.1 : Recommneded support spacing of copper pipes Pipe sagging due to spacing too far apart support spacing Figure 4.4: Effect of improper support spacing 4-3

52 4.4 Ver ertical Runs Vertical pipe runs (of small sizes up to 5/8") are usually mounted on walls by nailing them with wall clips. This is an easy and quick method of installation. An alternative method is by using pipe brackets. Simple saddle brackets made with angle iron are mounted with wall plugs onto the wall. The pipes are then clamped onto these brackets. This method is especially good for heavier and larger pipe sizes. TOP VIEW bracket copper pipe Figure 4.5: Vertical pipe installation on saddle brackets 4-4

53 Another method of running these pipes is by using electrical cable trays. These pre-fabricated trays are mounted onto the wall by using saddle brackets. The copper pipes are then clamped onto the trays. The main advantage of using these trays is a very neat, organized, and clean installation. Figure 4.6: Vertical pipe installation on cable tray Another similar method is by using trunking boxes. The trunking can be mounted straight onto the wall with wall plugs or on brackets. Copper pipes are then inserted into the trunking. The main advantage is that the pipes will be covered and protected from damage. Figure 4.7: Trunking box 4-5

54 In some instances, the pipes are required to go through a floor slab. A suitably sized hole must be made in the floor (e.g. by coring method) to accommodate all the pipes going through. Suitable brackets may then be fabricated to hold the pipes together. brackets core hole floor slab Figure 4.8: Pipe run through floor slab 4-6

55 4.5 Pipe Bends Copper pipes MUST NOT be bent with bare hands. This will cause the pipe to dent or collapse at the bent area. Use the proper pipe bending tool and the correct tool size corresponding to the required pipe diameter. Figure 4.9: Pipe bender Pipes up to ¾ can be bent by using the pipe bending tool. Generally, larger pipes are not bent but rather copper elbows are used instead. The elbows are brazed onto straight lengths of pipe. Figure 4.10 : Copper elbow An application example will be making an oil trap. The pipe bending tool is used to bend the two U-shape of the trap. For the larger pipe size, braze together 4 elbows to form the trap. Figure 4.11 : Example of how elbows can be used to create U-traps. 4-7

56 It is very common to find refrigerant pipe runs having to go over obstacles, e.g. concrete beams and columns, existing pipe works, and electrical conduits. To do this, many elbows and bends are used along the way. As much as this is necessary, it is important to keep within the specified maximum quantity of bends for the system. Refer to Table Vibr ibration and Noise in Piping Improper pipe installation may create undesirable vibration and noise. The effect of such vibrations are: a) Physical damage to the piping, mainly due to fatigue failure along the brazed joints. This lead to loss of refrigerant and subsequent compressor damage. b) Transmission of noise along the pipe into occupied spaces. The vibrations along the pipes are generated by the rigid connection of the piping to the compressor. It is impossible to eliminate vibration in piping, it is only possible to mitigate its effect. The indoor and outdoor air conditioning units have the internal piping designed to give minimal vibrations at the point of pipe connection. Thus, it is vital that the external piping must be designed and run properly to prevent unnecessary excessive vibrations. Several points for consideration: 1) In general, pipe vibration can be reduced by having flexibility in the piping and using isolation type hangers. Do not clamp the pipe too near to the outdoor unit (which houses the compressor) as this will increase the pipe rigidity at the connection. Allow sufficient length before putting in the first clamp or pipe support. 2) Vibration and noise radiation from a piping system may also be caused by gas pulsations due to the compressor action or from turbulence of high velocity refrigerant flow in the pipes. This is more apparent along the hot discharge line, e.g. during heating mode. Noise resulting from gas pulsations is usually objectionable only when the piping characteristics of the system result in amplification of the pulsation due to resonance. Such problems may be reduced by changing the size and length of the resonating pipe. Mufflers may also be added. Turbulence noises may be overcome by using a larger pipe to reduce the refrigerant velocities. 3) When the pipes penetrate through walls or floors, provide sufficient clearance to prevent vibration contact of the pipe surface with the hole. 4) Flexible metal hose may be used to absorb vibration transmitted along smaller pipe lines. These should be installed at right angles to the direction of vibration for best effect. However, such metal hose is not suitable for larger pipes because it is not actually flexible unless the ratio of length to the diameter is relatively great. Since, in practice, the length which can be used is often limited, it follows that flexibility is reduced with larger pipe size. 4-8

57 Appendix A Common Compressor Failure There are a few possibilities of compressor failure due to long piping installation. These are: a) Liquid flood back. b) Flooded starts. c) Liquid slugging. d) Loss of lubrication. e) Over heating. f) Contamination. g) Presence of moisture. h) Refrigerant migration. i) Liquid compression mechanism. A.1 Liquid Floodbac loodback Liquid floodback can be termed as the continuous return of liquid refrigerant (instead of vapour) to the compressor during the running cycle. Typical common causes of liquid floodback are: a) Over-charge of refrigerant. b) Return air filter dirty. c) Dirty coil. d) Return air duct too small. e) Evaporator blower dirty. f) Evaporator blower motor faulty. g) TXV or capillary tube oversized. h) Superheat setting is too low. With long piping installations, the main contributing factor is the extra refrigerant charged to the system, which causes this liquid floodback phenomenon. Liquid floodback will cause dilution of the compressor oil and also will wash oil from the moving parts surfaces. This in turn will lead to overheating of the mechanical bearing surfaces as the lubricating properties of the oil deteriorate and friction builds-up. Due to such causes, common compressor components affected are: a) Broken valves (suction and discharge). b) Seized bearings (main and cage). c) Seized connecting rods and pistons. d) Burnt motor due to mechanical fragments or high current draw. To prevent such undesirable situation, there are several ways to prevent it: a) Adequately sized accumulator. b) Ensure sufficient superheat at the suction. c) Correct size of expansion valve. d) Proper air flow/distribution. e) Proper refrigerant charge. A - 1

58 A.2 Flooded Starts ts Flooded starts are the result of refrigerant migration into the oil in the crankcase. This mixture causes foaming of the oil during start-ups. The foaming mixture gets sucked into the compression chamber and causes damage to the moving parts. Furthermore, during this period the oil level in the crankcase may reduce below the safety limit, causing insufficient lubrication. This phenomenon is more serious at lower outdoor temperatures as the refrigerant migration rate is higher. Generally, this absorption takes place during shutdown of the system. With such phenomenon, the affected compressor components are: a) Broken valves (suction and discharge IMMEDIATE). b) Blown gaskets IMMEDIATE. c) Loss of lubrication GRADUAL. With long piping installations, because of the higher refrigerant charge, the migration rate into the oil becomes higher. Therefore, during start-ups the foaming becomes more violent, subjecting the compressor to higher operating stresses. There are ways to minimise such undesirable situation. They are: a) Proper refrigerant charge. b) Correct amount of oil in the crankcase (refer to manufacturer specifications and data sheets). c) Install crankcase heater. d) Pump down cycle. A.3 Liquid Slugging ging This is a term to describe a compressor pumping liquid refrigerant, oil, or both. In other words, it is liquid compression. This is characterised by a loud metallic clatter within the compressor, accompanied by extreme vibrations. Liquid slugging is a severe form of liquid floodback. Slugging normally appears at start up when liquid refrigerant has migrated to the sump. Liquid slugging occurs due to: a) No crankcase heater fitted. b) Defective crankcase heater or not connected. c) Compressor experiencing liquid floodback (see Section A.1: Liquid floodback). d) Overcharge of refrigerant. e) Overcharge of oil in crankcase. Again the main contributing factor with long piping installations is the extra refrigerant charge required by the system. If not careful, this may cause large amount of liquid refrigerant flooding into the compressor. Typical failures related to this phenomenon are: a) Damaged piston, connecting rod, crankshaft and scroll orbits. b) Suction or discharge reed broken. c) Motor damaged due to broken internal components. There are several ways to prevent liquid slugging: a) Pump down control system. b) Crankcase heater must be energised at least 24 hours before the initial start-up. c) Adequate accumulator sizing. d) Proper superheat setting on the expansion valve. e) Correct size of expansion valve. f) Proper refrigerant charge. A - 2

59 A.4 Loss of Lubrication Introduction In any refrigeration system, oil and refrigerant are always present. The main purpose of oil is to lubricate the mechanical moving parts of the compressor. Liquid refrigerant and oil are miscible in one another and their magnitude of miscibility will depend on the type of refrigerant, the temperature, and the pressure which both are exposed to. It is because of this miscibility that a certain amount of oil will always leave the compressor s crankcase and be circulated with the refrigerant. Although oil is always treated as a lubricant to reduce mechanical wear and friction, oil actually accomplishes many more purposes. The other functions of oil are: a) Act as a seal between the discharge and suction sides of the compressor. b) Act as a noise dampener by reducing internal mechanical noise within a compressor. c) Performs heat transfer task by sweeping away any heat from internal rotating and stationary parts. Causes and prevention Loss of lubrication is defined as the absence or lack of oil in the crankcase. Generally, this will occur when the rate of oil return is lower than the rate it is pumped out of the compressor. The system must allow oil to return to the compressor at the rate it leaves; else it can cause overheating problems. The common causes of loss of lubrication are: a) Low refrigerant velocity (e.g. due to wrong pipe sizing). b) Insufficient or no oil traps. c) Very frequent ON/OFF cycling of the compressor. d) Low loads which reduces the refrigerant flow rate. e) Liquid flood back. f) Oil trapped in the system. g) Loss of refrigerant charge. With long piping installations, there are several points to be aware of: 1) The higher refrigerant charge required in the system may dilute the oil in the compressor, causing deterioration of the lubricating properties. See Section 2.1: Additional oil. 2) The long piping may cause the velocity of the refrigerant in the pipe to reduce, due to friction forces. As a result, the oil becomes more difficult to be carried along by the refrigerant. Hence, the rate of oil return to the compressor reduces. To prevent oil loss, the following choices of solutions can be considered: a) Apply compressor minimum run time setting to ensure oil return after start up. This will allow time for the compressor to return the oil from the system. Too frequent start-stop cycles will reduce this run time and can cause oil return problems. b) Correctly size the pipe diameter length, and reduce the number of bends. Where necessary install proper oil traps on the vertical pipe lines. c) Do not overcharge the system to prevent oil dilution. A - 3

60 A.5 Overhea erheating Compressors generate heat through compression, motor windings, and friction at load bearing surfaces. It is this heat that causes the compressor external shell and discharge port to be hot. Compressors are designed to withstand this high temperature up to a specified limit. In most common applications, the highest operating temperature allowed at the discharge line is 135 o C (275 o F) for reciprocating compressors and 115 o C for rotary compressors. At temperatures higher than this limit, the lubricating properties of the oil will deteriorate. The motor winding insulation will also begin to breakdown, causing damage to the compressor. In some instances, some of the moving parts may also seize together. The causes of over heating can be categorized into three broad areas: a) Refrigeration i) Improper setting of controls (TXV, pressure regulators, hot gas bypass, pressure control switches, etc) causing insufficient refrigerant flow through the compressor, reducing the motor cooling ii) Lack of proper suction line insulation causing a higher return gas temperature to the compressor iii) Low suction pressure due to undersized evaporator and loss of refrigerant which causes a lower refrigerant flow rate and reducing the motor cooling iv) High discharge pressure due to blocked condenser, insufficient air circulation, recirculation of hot discharge air, undersized discharge line, condenser fan motor failure and refrigerant overcharge. v) Highly superheated return gas temperature b) High compression ratio Compression ratio is defined as the ratio of the compressor discharge pressure to the suction pressure. High ratios generally occur when the outdoor ambient temperature becomes very high while trying to maintain a cool indoor temperature. It is important that the system operates within the specified operating temperature limits to prevent such high ratios. c) Electrical i) Voltage unbalance between phases, causing excessive winding temperature. ii) Current unbalance between phases, causing excessive winding temperature. iii) Single phasing of the power supply, causing high winding temperature. iv) Supply voltage too high (e.g. > 15% of nominal voltage). v) Faulty capacitors and contactors. vi) Rapid ON/OFF cycling of the compressor. With the long piping installation, the main contributing factor which may lead to compressor overheating is due to insufficient oil return and oil dilution. This lack of lubrication will cause friction to build up in the moving parts and cause the temperature to rise. To prevent overheating: a) Maintain the suction and discharge pressure at safe levels. b) Control the return gas temperature by i) Insulating the suction line. ii) Setting adequate superheat. c) Ensure sufficient lubrication (See Section A.4: Loss of lubriation). A - 4

61 A.6 Contamination tion Contamination is the presence of foreign substances in the refrigerant system. Some foreign matter can cause chemical reaction or change the chemical composition of material within the system. There are several types of contamination: a) Acid in system from previous compressor change. b) Flux from solder joints. c) Copper shavings. d) Water. e) Dirt. f) Air. The effects of contamination are: a) Blocked oil passages leads to bearing failure. b) Motor failures due to solid shorting windings. c) High head pressure due non condensable gases. d) Moisture in the system - forms acid in the system which attacks the metal and windings. Moisture will also cause the expansion device to freeze-up internally. The longer the piping, the higher the chances that contamination will get into the system. There are several ways to eliminate such undesirable situations: a) Air Evacuate the system thoroughly before charging. b) Moisture Evacuate the system thoroughly before charging. c) Foreign matter Apply care to workmanship. Use filter-driers. Presence of Moisture Of all the contaminants, moisture in a HVAC system is the most harmful. Moisture will reduce the life span of the HVAC system. The possible causes for moisture to be present are: a) Open system exposed to air and moisture. b) Compressor tubes left open. c) Leak in system (particularly on the low side). d) Wet rags or water to cool poor solder joints. e) Wet refrigerant. f) Lack of knowledge with the use of hygroscopic oils. g) Incorrect evacuation process. Again, with long piping installations, the chances are higher that moisture may enter into the refrigeration system. This is due to the extra brazing/welding of the long pipe sections. Figure A.1 shows the progression of compressor failure due to the contamination of foreign substances, while Figure A.2 shows the same when the contamination is due to air and moisture. A - 5

62 Entry of foreign substances. Entry into the compressor. Restriction of expansion valve, capillary tube & drier. Large pressure loss. Bearing seizure, locking between cylinder and piston. Valve breakage. Seating in valve Evaporator pressure drop. Large compression ratio. Lock. Overload or lock. Poor start. Insufficient cooling. Compressor overheating. Figure A.1: Causes and effects of foreign substances entry into refigerantion system. A - 6

63 Air entry. Moisture entry. Abnormal heating. Rise of the condenser pressure. Condensation to moisture. Hydrolysis with the refrigerant. Sludge deposit at the valve parts. Rise of the condensing temperature. Large compressor load. Acid generation. Sludge deposit. Icing of expansion valve & capillary tube. Bearing overheating. Motor overheating. Defective valve operation. Motor overheating. Motor insulation damage. Copper plating of rotating and sliding parts. Lubricating oil deterioration. Clogging of expansion valve & capillary tube. Lock. Motor burnout. Lack of cooling capacity. Wear & seizure of rotating and sliding parts. Wasteful power consumption. Motor burnout. Lock. Seizure of rotating and sliding parts. Insufficient cooling. Compressor overheating. Figure A.2 : Causes and effects of air and moisture entry to refrigerant system. A - 7

64 A.7 Refrig efriger erant Migration Refrigerant migration is described as the absorption of liquid refrigerant into the compressor oil during periods when the compressor is not operating for a long period. It occurs when the compressor sump becomes colder than the indoor coil. As a result of this temperature differential, there will be partial pressure differential of the refrigerant between these two locations. This will force the refrigerant to migrate to the compressor where it is absorbed into the oil. Although this type of migration is most pronounced in colder weather, it can also occur at relatively high ambient temperatures with remote type condensing units for air conditioning and heat pump applications. When the compressor is started in this condition, the following negative effects occur: a) During starting, the rapid pressure drop inside the compressor causes the refrigerant in the lubricating oil to explode out, thereby creating foaming in the sump. The foam, which contains lubricating oil is then sucked into the cylinder causing liquid compression. b) Liquid compression occurs because the liquid refrigerant and lubricating oil in the cylinder are compressed. This will damage the valve parts and bearings. c) As the foam gets pumped out of the compressor, the amount of lubricating oil inside the compressor is reduced. Thus, the bearings will not be lubricated sufficiently and seizure might occur. Long piping will require more refrigerant charge, therefore causing the migration rate to increase. The oil dilution also becomes more severe. All this gives potential problems to the compressor if care is not taken during installation and operation. The dangers of refrigerant migration can be prevented by having proper control of the refrigerant charge amount and by installing a crankcase heater. See Section A.2: Flooded starts and Section A.3: Liquid slugging. A - 8

65 Appendix B Suction Accumulator As mentioned in Section 2.4: Suction Accumulator, it is not recommended to install an additional suction accumulator in the long piping system as long as the maximum pipe length limits are not exceeded. All heat pump products have built in suction accumulator. Nevertheless, the following provides some information on the function and construction of an accumulator. The accumulator is a vessel which serves as a storage container for liquid refrigerant which has flooded through the system. It has a provision for metering the return of oil and liquid to the compressor at a rate in which the compressor can safely operate. Each accumulator features an inlet deflector that bends refrigerant flow to prevent internal splashing and aid in the collection of refrigerant oil in the bottom of the unit. A U-tube is connected to the outlet connection of the accumulator. As refrigerant gas leaves the accumulator, oil is pulled through an oil return orifice and returned to the compressor. Solid copper fittings allow for easy installation. Suction accumulator assures only the return of refrigerant vapour and prevents compressor failure due to liquid refrigerant entering the compressor. Figure B.1: Refrigerant flow inside accumulator. B - 1

66 Oil return orifice Figure B.2: Internal layout of suction accumulator. B - 2

67 B.1.1 Applica pplications Refrigerant flood back in a system is one of the most common causes of compressor failure. Excessive liquid refrigerant dilutes the oil in the compressor crankcase causing wear and tear to the moving parts. Complete loss of oil in the compressor can result in broken rods and crankshafts. In heat pump systems, an accumulator can act as a receiver during the heating cycle, when the system load imbalance results in excessive liquid refrigerant in the system. Flooding can occur in a heat pump system whenever the cycle is switched between cooling and heating as there may be liquid that has not vaporized. The liquid is then pumped back to the compressor. This may also occur during the defrost cycle, where the liquid has not cleared the evaporator on start up or termination of the defrost cycle, or during low ambient heating cycle when there is insufficient air temperature to vaporize the liquid. All the above mentioned problems will be compounded with long piping installations due to the additional refrigerant charge. The accumulator should be located in the compressor s common suction line between : a) The reversing valve and compressor in heat pump units, and b) The evaporator and compressor in cooling only units. It must also have provisions for a positive return of oil to the compressor so that oil does not become trapped in the accumulator. The liquid refrigerant and oil must be metered back to the compressor at a controlled rate to avoid damage to the compressor. Therefore, proper sizing of the oil orifice is required. The actual refrigerant holding capacity needed for an accumulator is also determined by the particular application and should be selected to hold the maximum liquid refrigerant flood back anticipated. B.2 Installation tion Install the accumulator in the common suction line as close to the compressor as possible. Be sure that the inlet connection is connected to the common suction line and the outlet connection goes to the compressor. Always install accumulators in the vertical position. When the compressor-condensing unit is located indoors, there may be a problem of suction accumulators that sweat and drip. It is necessary to completely insulate the accumulator to be vapour sealed to prevent condensation forming under the insulation. A rusting problem may occur if the accumulator is exposed to moisture for long periods of time. Care must be taken to prevent the paint from being burnt off during the welding process to avoid the metal from being exposed. When the compressor is being changed due to severe compressor burnout, the suction accumulator should also be changed. The contaminants and particles that are caught in the accumulator during the burnout can return to the new compressor and cause damage. It is also possible that oil from the first compressor may be stored in the accumulator and the excess oil return may cause failure. B - 3

68 B.3.3 Sizing and Selecting an Accumula ulator Suction accumulators should never be selected based on connection sizes only. It is more important to select an accumulator based on the minimum pressure drop, proper oil return, and the amount of refrigerant it is required to hold. Suction accumulators are meant to assist with momentary flooding and migration. However, under severe conditions the accumulator must have sufficient volume to prevent over flowing and causing damage to the compressor. As a guideline, the accumulator must have adequate liquid holding capacity of not less than 50% of the entire system charge. The accumulator should not add excessive pressure drop to the system. A properly sized oil return orifice ensures positive oil return to the compressor. The recommended orifice size is 1 mm. B - 4

69 Appendix C Oil Seperator There is no requirement to install additional oil separators in the system if the pipe length limits are not exceeded, as described in Section 2.5 : Oil Separator. The following provides some additional information about the function, construction, and installation of commercially available oil separators. C.1.1 Introduction Refrigeration compressors are lubricated by refrigerant oil that circulates from the crankcase or housing. When the compressor operates, refrigerant oil will leave the compressor in a mixture with the hot compressed refrigerant gas. Small amounts of oil circulating through the system will not affect the system s performance. However, too much circulating oil interferes with the operation of flow controls, evaporator, condenser, and filter driers. At low temperature installations, refrigerant oil thickens and becomes difficult to move out of the evaporator. Accumulation of refrigerant oil in the evaporator would affect evaporator efficiency leading to compressor failure. No matter what type of oil separator, they are not 100% efficient; some small quantities of oil will continue to be transported with the discharge gas and refrigerant through the system. Placing an oil separator between the compressor discharge and the condenser will protect the refrigeration system. The oil separator will maintain correct oil level in the compressor, reduce oil trapping, and improve on system reliability. Figure C.1: Oil sperator within a cooling system circuit. C - 1

70 Figure C.2 : Internal layout of oil seperator C.2 Function A mixture of refrigerant and oil from compressor enters into the inlet of the oil separator. This mixture flows through a screen and baffle arrangement to cause the fine particles of oil to gather and drop to the bottom of the oil separator. The refrigerant gas passes through the outlet screen to trap residual oil particles, and passes oil free to the condenser. The refrigerant oil gathers in the bottom of the oil separator unit, where a float operated needle valve opens to allow the return of oil to the compressor. Oil returns quickly to the compressor because of the higher pressure in the oil separator than in the crankcase. When the oil level has lowered, the needle valve will reseat to allow oil to build-up again in the separator. C - 2

71 C.3.3 Installation tion The oil separator must be primed with the correct type and grade of compressor oil. It must be mounted securely in a vertical position. If space permits, the separator can be installed inside the unit, else, it can be installed externally. The oil separator should be installed in the discharge line as close as possible to the compressor. An initial charge of refrigerant oil must be added to the oil separator to actuate the float mechanism to return oil to the compressor. Use the same type of oil that is in the crankcase of the compressor. The discharge line from the compressor is assembled to the inlet connection of the oil separator and a line is connected from the outlet connection of the oil separator to the inlet of the condenser. The smallest connection on the oil separator is the oil return connection and a line is run from this to the compressor crankcase or suction pipe line. To do this, it may be necessary to cut the existing internal piping of the unit and modify it. The pipe line from the separator to the condenser should be carried about 50mm higher than the condenser and pitched with a downward slope into the condenser inlet connection. In this way, should any condensation occur in this line at the condenser connection, it will drain forward into the condenser. The body of the oil separator should be insulated so that it retains some heat during the compressor idle periods. Otherwise, it may act as a primary condenser on start up. If this should occur, the separator will feed condensed liquid refrigerant back to the compressor crankcase, causing liquid hammering, oil dilution, and risk of more mechanical damage. This situation can easily and safely be eliminated by the addition of an electrical off cycle heater cable of low wattage, applied to the separator body below the insulation. C.4 Maintenance When the float valve jams, oil stops flowing to the compressor preventing sufficient oil return. A periodic inspection will help prevent such undesired situation. Oil separators stop working when solid materials such as oxide scale and carbon jam the float mechanism and block the orifice to the compressor. The high discharge temperatures of the compressor may cause solid particles to be formed in the oil. These particles will end up in the bottom of the separator, jamming the float mechanism and blocking the valve seat. If the oil return is continually hot, the oil float valve may be leaking, or it is being held open by sludge or foreign matter. The backpressure will be affected, reducing system capacity. A compressor that is pumping excessive oil will also cause the return line to be continually hot. If the oil return line is continually cold, there may be condensation of liquid refrigerant in the oil separator. This liquid, when entering the compressor crankcase could cause lubrication failure within the compressor. This should not be allowed to happen. When piping up long discharge lines, vertical runs of piping should include oil traps every 3 meters of rise to prevent excess oil in the discharge line from returning to the oil separator during the off cycle. During long off-cycles or long manual shutdowns, liquid refrigerant may collect in the oil separator. The return of liquid refrigerant to the compressor through the oil return line may cause slugging and possible damage to the compressor. A check valve installed on the outlet line of the oil separator will help prevent the liquid refrigerant from returning to the compressor. Insulating the oil separator will prevent it from acting as a condenser and passing heat to the surrounding air. The addition of a filter in the oil return line will help keep the oil clean. C - 3

72 Appendix D Useful Tables and Charts Nominal pipe or tube size (in.) Smooth bend elbows Smooth Bend Tees 90 o Flow- Straight-Thru Flow 90 o Std 1 90 o 45 o 180 o Long 45 o Std 1 Thru No Reduce Reduce Street 1 Street 1 Std 1 Radius 2 Branch Reduction d 1/4 d 1/2 3/ / / / / / / NIL 10 NIL NIL 13 NIL NIL 16 NIL NIL 18 NIL NIL 20 NIL NIL 23 NIL NIL 26 NIL NIL 30 NIL Note : 1) R/D approximately equal to 1. 2) R/D approximately equal to 1.5. Table D.1: Various bends losses equivalent length D - 1

73 Nominal Pipe or Tube Size (in.) Mitre Elbows 90 o Ell 60 o Ell 45 o Ell 30 o Ell 3/ / / / / / / Table D.2: Various elbow losses equivalent length D - 2

74 Nominal pipe or tube size (in.) Valve Losses in Equivalent feet of pipe 1 Globe 2 60 o - Y 45 o - Y Angle 2 Gate 5 Swing Check 3 Y - Type Strainer 6 Lift Check Flanged End Screwed End 3/ NIL NIL 1/ NIL 3 3/ NIL NIL 5 1 1/ NIL 9 1 1/ NIL / / NIL NIL NIL NIL NIL NIL NIL NIL NIL NIL NIL NIL NIL NIL NIL NIL NIL Globe & Vertical Lift Same as Globe Valve 4 Angle Lift Same as Angle Valve Note : 1) Losses are for all valves in fully open position and strainers clean. 2) These losses do not apply to valves with needle point type seats. 3) Losses also apply to the in line ball type check valve. 4) For Y pattern globe lift check valve with seat approximately equal to the nominal pipe diameter, use values of 60o Y valve for loss. 5) Regular and short pattern plug cock valves, when fully open, have same loss as gate valve. For valve losses of shoty pattern plug cocks above 6 inches check manufacturer. 6) For thru 3/16 inch perforations with screen 50% clogged, loss is double. Table D.3: Valve losses equivalent length D - 3

75 Nominal pipe or tube size (in.) Sudden Enlargement* d/d Sudden Contraction* d/d Sharp Edge* Pipe Projection* 1/4 1/2 3/4 1/4 1/2 3/4 Entrance Exit Entrance Exit 3/ / / / / / / NIL NIL NIL NIL NIL NIL NIL NIL 16 NIL NIL NIL NIL 18 NIL NIL NIL NIL 20 NIL NIL NIL NIL NIL NIL NIL NIL NIL NIL NIL NIL NIL NIL * Enter table for losses at smallest diameter d. Table D.4: Special fitting losses equivalent length D - 4

76 Classification Nom. Tube Size (in.) OD (in.) Stubbs Gage t w (in.) ID (in.) Transverse area (sq in.) Minimum Test Pressure (psi) Weight of Tube (lb/ft) WT of Water in Tube* (lb/ft) Outside Surface (sq ft/ft) HARD 1/4 3/ /8 1/ /2 5/ /4 7/ / /4 1 3/ /2 1 5/ Govt. Type "M" 2 2 1/ lb Working 2 1/2 2 5/ Pressure 3 3 1/ /2 3 5/ / / /8 NIL /8 NIL HARD 3/8 1/ /2 5/8 NIL /4 7/8 NIL /8 NIL /4 1 3/8 NIL /2 1 5/8 NIL Govt. Type "L" 2 2 1/8 NIL lb Working 2 1/2 2 5/8 NIL Pressure 3 3 1/8 NIL /2 3 5/8 NIL /8 NIL /8 NIL /8 NIL HARD 1/4 3/ /8 1/ /2 5/ /4 7/ / /4 1 3/ Govt. Type "K" 1 1/2 1 5/ lb Working 2 2 1/ Pressure 2 1/2 2 5/ / /2 3 5/ / /8 NIL /8 NIL SOFT 1/4 3/ /8 1/ /2 5/ /4 7/ / /4 1 3/ /2 1 5/ / /2 2 5/ / /2 3 5/ / /8 NIL /8 NIL * To change Wt of Water in Tube (lb/ft) to Gallons of Water (gal/ft), divide values in tableby Table D.5: Properties of Copper tube D - 5

77 Service R12, R22 AND R500 Chilled Water Condenser or Make Up Water Drain or Condensate Lines Steam or Condensate Hot Water Suction Line Liquid Line Hot Gas Line Pipe Hard copper tubing, Type L* Steel pipe, standard wall Lap welded or seamless for sizes larger than 2" IPS. Hard copper tubing, Type L* Steel pipe Extra strong wall for sizes 1 1/2" IPS and smaller. Standard wall for sizes larger than 1 1/2" IPS. Lap welded or seamless for larger than 2" IPS. Hard copper tubing, Type L* Steel pipe, standard wall Lap welded or seamless for larger than 2" IPS. Black or galvanized steel pipe** Hard copper tubing** Galvanized steel pipe** Hard copper tubing** Galvanized steel pipe** Hard copper tubing** Black steel pipe** Hard copper tubing** Black steel pipe Hard copper tubing** Fittings Wrought copper, wrought brass or tinned cast brass. 150lb welding or threaded malleable iron. Wrought copper, wrought brass or tinned cast brass. 300lb welding or threaded malleable iron. Wrought copper, wrought brass or tinned cast brass. 300lb welding or threaded malleable iron. Welding, galvanized, cast, malleable or black iron. *** Cast brass, wrought copper or wrought brass. Welding, galvanized, cast, malleable or iron. *** Cast brass, wrought copper or wrought brass. Galvanized, drainage, cast or malleable iron.*** Cast brass, wrought copper or wrought brass. Welding or cast iron.*** Cast brass, wrought copper or wrought brass. Welding or cast iron.*** Cast brass, wrought copper or wrought brass. * Except for sizes 1/4" and 3/8" OD where wall thicknesses of 0.30 and 0.32 inch are required. Soft copper refrigerantion tubing may be used for sizes 1 3/8" OD and smaller. Mechanical joints must not be used with soft copper tubing in sizes larger than 7/8" OD. ** Normally standard wall steel pipe or Type M hard copper tubing is satisfactory for air conditioning applications. However, the piping material selected should be checked for the design temperature-pressure ratings. *** Normally 125lb cast iron and 150lb malleable iron fittings are satisfactory for the usual air conditioning applications. However, the fitting material selected should be checked for the design temperature-pressure ratings. Table D.6: Pipe type recommendations D - 6

78 Copeland s Summit Series Compressor Specification Model Current (A) Operating Conditions CCH (W) LRA RLA P T max ( o max (psig) C) ZR36K3-PFJ-501 ZR42K3-PFJ-501 ZR47KC-TFD ZR61KC-TFD-501 ZR72KC-TFD-501 ZR36K3E-PFJ ZR47KCE-TFD ZR61KCE-TFD ZR72KCE-TFD Table D.7: Compressor s operating specifications Model ZR36K3-PFJ-501 ZR42K3-PFJ-501 ZR47KC-TFD-501 ZR61KC-TFD-501 ZR72KC-TFD-501 ZR36K3E-PFJ-501 ZR47KCE-TFD-501 ZR61KCE-TFD-501 ZR72KCE-TFD-501 White Oil Charge MMMA Oil Refrigerant (CC) Charge (CC) charge Initial Refill Initial Refill kg lb N/A N/A N/A N/A N/A N/A N/A N/A N/A N/A N/A N/A N/A N/A N/A N/A N/A N/A Table D.8: Compressor s charging specifications D - 7

79 Matsushita s Compressor Specification Model 2PS164D2BC02 2KS340D3AA02 2JS464D3AA02 2JS35C225ASA 4 2J44C3R225A 4 2JS442P3AA01 5 2JS356P3AA01 5 2KS224D3AC02 2JS350D3BB02 2JS438D3AA02 5CS102XEB 2RS127D5BB02 5RS092XAB 2KS210D3BB02 2JS350D3BA02 2JS438D3BA02 2JS464D3BC02 5RS080DRSM329 5PS102DPSM370 5PS132DPSM371 2PV164N7BB02 6 2JS324D3AB07 2JS438D3JA02 2KS206D3AB04 2KS28S236A6F 7 2J35S236AB 7 4PS164DAA 4KS225DAA 4JS350DAC 4JS435DAC 4JS435PAA 5 4JS350PAA 5 4PS132DAA Current (A) Operating Conditions Oil Charge (cc) Refrigerant LRA RLA P max (MPa) T max ( o C) Initial Type Charge (kg) 18 / / 3.9 N / A A / / / / A A N / A A N / A A N / A A N / A A / / 5.90 N / A A / / 10.5 N / A A / / 13.1 N / A A N / A N / A B / / A N / A N / A B / / A / / A / / A / / A / / B / / B / / B N / A 7.4 N / A A / / A / / A / / A A A N / A N / A N / A N / A 350 B N / A N / A N / A N / A 410 B N / A N / A N / A N / A 700 B N / A N / A N / A N / A 700 B B B / / B Note: 1) Oil Type A - ATMOS M60 or SUNISO 4GDID. 2) Oil Type B - RB68A or FREOL ALPHA 68M. 3) All current is based on 220V /240V respectively unless stated otherwise. 4) 240V supply voltage. 5) 380V supply voltage. 6) 110V supply voltage. 7) 230V supply voltage. 8) Pa = PSI x Table D.9: Compressor s operating specifications D - 8

80 Toshiba s Compressor Specification Model Current (A) Operating Conditions Oil Charge (cc) Refrigerant LRA RLA P max (MPa) T max ( o C) Initial Type Charge (kg) PH120X1C - 4DH / / A PH165X1C - 4DH / / A PH160X1C - 4DZ / / A PH135X1C - 4DZ / / A Note: 1) Oil Type A - SUNISO 4GSD. 2) Pa = PSI x ) All current is based on 220V /240V respectively unless stated otherwise. 4) kg = lb x Table D.10: Compressor s operating specifications D - 9

81 Bristol s Compressor Specification Model Operating Current (A) Oil Charge (fl.oz) Refrigerant Conditions CCH (W) Charge (l b) LRA RLA P max (PSIG) Initial Refill Type H23B30QABK A H23B26QABC A H23A353DBE A H23A383ABK A H23A383DBE A H23A463DBE A H23A623DBE A H23B24QABK A H2NG204DRE N / A B H2NG294DPE N / A B H23B32QABK A H23B35QABK A H23A383ABC A H23A463ABK A H25A62QCBC A H25A62QDBL A H25A62QDBL A H25G094DBD 4 & N / A C H25G124DBD 2 & N / A C H25G144DBD 2 & N / A C H2NG204DRE N / A B H2NG204DRE N / A B H2NG244DRE N / A B H2NG244DRE N / A B H2NG294DPE N / A B H2NG294DPE N / A B H25G104DBD N / A C H25G184DPD N / A C H25G204DPD N / A C H25G244DPD N / A C H25G294DPD N / A C H25G094DBE N / A C H25G104DBE N / A C H25G124DBE N / A C Note: 1) 220V / 240V / 50Hz supply voltage. 2) 230V / 208V / 60Hz supply voltage. 3) 380V / 415V / 50Hz supply voltage. 4) 230V / 200V / 60Hz supply voltage. 5) 220V / 200V / 50Hz supply voltage. 6) 460V / 60Hz supply voltage. 7) 380V / 460V / 60Hz supply voltage. 8) Oil Type A - Specification ) Oil Type B - Zerol 150T. 10) Oil Type C - Specification ) Pa = PSI x ) L = fl.oz x ) kg = lb x Table D.11: Compressor s operating specifications D - 10

82 Model Operating Current (A) Oil Charge (fl.oz) Conditions LRA RLA P max (PSIG) Initial Refill Type CCH (W) Refrigerant Charge (l b) H25G144DBE N / A C H25G184DPE N / A C H25G204DPE N / A C H25G244DPE N / A C H25G294DPE N / A C H23B30QABC A H26A72QDBE A H26A72QDBL A H2NG184DPE N / A B H2NG184DPD N / A B H23B20QABC A H23B24QABK A Note: 1) 220V / 240V / 50Hz supply voltage. 2) 230V / 208V / 60Hz supply voltage. 3) 380V / 415V / 50Hz supply voltage. 4) 230V / 200V / 60Hz supply voltage. 5) 220V / 200V / 50Hz supply voltage. 6) 460V / 60Hz supply voltage. 7) 380V / 460V / 60Hz supply voltage. 8) Oil Type A - Specification ) Oil Type B - Zerol 150T. 10) Oil Type C - Specification ) Pa = PSI x ) L = fl.oz x ) kg = lb x Table D.12: Compressor s operating specifications D - 11

83 Oil Dilution Ratio As discuss at Section 2.1, liquid flood back will cause dilution of the compressor oil. This dilution process will deteriote the lubricating properties of oil. The formula shown below is to help end users to select a proper refrigerant charge and oil charge: R = m oil m oil + m refrigerant x 100%...(A.1) R = (rv) oil (rv) oil + m refrigerant x 100%...(A.2) R = 0.9*oil(cc) x 100% 0.9*oil(cc) + R22 Weight(g)...(A.3) Where: 1) R is a dilution ratio and must be greater than 22%. 2) Volume for oil in centimeter cube (cc). 3) ρ oil is assume to be 0.9g/cc. Sample calculation A-01: a) Standard factory testing length = 7.6m / 24.9 ft b) Actual pipe length = 10.0m / 32.8ft c) Compressor specifications: i) Initial oil charge = 38fl.oz. (Brand new compressor) ii) Refill oil charge = 34fl.oz. (Not brand new compressor) d) Standard factory charge = 2.50kg Extra length = 32.8ft 24.9ft = 7.9 ft With 10ft = 3fl.oz., therefore extra oil charge for extra length of: 7.9ft = 2.37fl.oz. 1 fl.oz» 30cm 3» 0.03l Total volume of oil = 38.00fl.oz fl.oz = 40.37fl.oz = cm 3 Substitute values to equattion A.3 : R = = = 0.9*oil(cc) x 100% 0.9*oil(cc) + R22 Weight(g) 0.9 x x 100% 0.9 x % R is greater than 22%, therefore criteria met. D - 12

84 ALCO Oil Separator Capacity Ratings in TONS Model Number At Evaporator Temperature R12 R22 R502 R134A Flanged Sealed -40 o F +40 o F -40 o F +40 o F -40 o F +40 o F -40 o F +40 o F A-F A-W A-F A-W A-F A-W A-F A-W A-F A-W A-F A-W NIL A-W A-F A-W A-F A-W Table D.13: Oil separator capacity ratings in TONS ALCO Oil Separator Capacity Ratings in KWS Model Number At Evaporator Temperature R12 R22 R502 R134A Flanged Sealed -40 o C +40 o C -40 o C +40 o C -40 o C +40 o C -40 o C +40 o C A-F A-W A-F A-W A-F A-W A-F A-W A-F A-W A-F A-W NIL A-W A-F A-W A-F A-W Table D.14: Oil separator capacity ratings in KWS D - 13

85 ALCO Oil Separator Dimensional Data Sealed Type (Style No. 1) Dimensions Connection Size Model No. Style No. A in mm in mm in A-W / A-W / A-W / A-W / A-W / A-W / A-W / A-W / A-W / Table D.15: Oil separator dimensions for style 1 B mm ALCO Oil Separator Dimensional Data Sealed Type (Style No. 2) Dimensions Connection Size Model No. Style No. A B C in mm in mm in mm in mm A-W / A-W / Table D.16: Oil separator dimensions for style 2 ALCO Oil Separator Dimensional Data Sealed Type (Style No. 3) Dimensions Connection Size Model No. Style No. A B C in mm in mm in mm in mm A-F / A-F / A-F / A-F / A-F / A-F / A-F / A-F / Table D.17: Oil separator dimensions for style 3 Figure D.1 : Different styles of oil separator D - 14

86 Fitting Size Shell Diameter Oil Separator Cross Reference Chart Sealed Type - Float Valve ALCO AC & R TEMPRITE Model Length Length Length Model Model inch mm inch mm inch mm 1/ S / S /2 A-W S /8 A-W S /8 4" A-W S /8 A-W S /8 A-W S /8 A-W / S / S /8 6" A-W S /8 A-W S /8 A-W S Table D.18: Oil separator (sealed type) cross reference chart Fitting Size Shell Diameter Oil Separator Cross Reference Chart Flanged Type - Float Valve ALCO AC & R TEMPRITE Model Length Length Length Model Model inch mm inch mm inch mm 1/2 A-F S /8 A-F S /8 A-F S /8 4" A-F S /8 A-F S /8 A-F /8 A-F * S " 2 1/8 A-F * S / S /8 8" S / / S " 3 1/ / S " 3 5/ /8 14" Table D.19: Oil separator (float type) cross reference chart D - 15

87 Capacity in Tons of Refrigeration Model Number R134A R404A/R o F -20 o F 0 o F +20 o F +40 o F -40 o F -20 o F 0 o F +20 o F +40 o F A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS Table D.20: Suction accumulator capacity selection chart Capacity in Tons of Refrigeration Model Number R22 R o F -20 o F 0 o F +20 o F +40 o F -40 o F -20 o F 0 o F +20 o F +40 o F A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS A-AS Table D.21: Suction accumulator capacity selection chart D - 16

88 Model Number ALCO Accumulator Dimensional Data A B C D Unit Weight Diameter Length (LBS) (in) (in) Fitting Size (Nominal) Fitting Separation (in) Tons R22 (+40 o F) Holding Capacity 40 o F Liquid R404A /R o F Liquid R502/R22/ R134A A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / A-AS / Table D.22: Suction accumulator holding capacity selection chart Figure D.2: Accumulator cross-section D - 17

89 Suction Accumulator Cross Reference ALCO VKMP PARKER Model OD Tube Fitting Nominal Length (Cap to Cap) Tons R22 (+40oF) Holding capacity (LBS)1 Model Fitting A-AS / VA-30-4S 1/ A-AS / VA-31-5S 5/ PA / A-AS / VA-32-5S 5/ A-AS / VA-32-6S 3/ A-AS / VA-35-5S 5/ PA / A-AS / VA-35-6S 3/ PA / A-AS / A-AS / A-AS /8 6 5/ A-AS / VA-44-5SRD 5/ PA C 5/ A-AS / VA-44-6SRD 3/ PA C 3/ A-AS / VA-54-6SRD 3/ PA C 3/ A-AS / VA-54-7SRD 7/ PA C 7/ A-AS / VA-56-6SRD 3/ PA / A-AS / VA-56-7SRD 7/ PA / A-AS / VA-57-7SRD 7/ PA / A-AS / VA-57-9SRD 1 1/ A-AS / VA-59-9SRD 1 1/ A-AS / VA-59-11SRD 1 3/ A-AS / A-AS / VA-610-7SRD 7/ A-AS / VA-610-9SRD 1 1/ A-AS / VA SRD 1 3/ A-AS / A-AS / VA SRD 1 5/ A-AS / Nominal Length (Cap to Cap) Tons R22 (+40oF) Holding capacity (LBS)1 Model Fitting Overall Length Tons R22 (+40oF) Holding capacity (LBS)3 Notes: 1) R22 at 40 o F evaporator. 2) R22. 3) R22 at 40 o F divided by ) R22 at 0 o F saturation. Table D.23: Suction accumulator cross reference chart D - 18

90 Suction Accumulator Cross Reference ALCO AC&R Refrigeration Research Model OD Tube Fitting Nominal Length (Cap to Cap) Tons R22 (+40oF) Holding capacity (LBS)1 Model Fitting A-AS / / A-AS / A-AS / / A-AS / A-AS / A-AS / A-AS / / A-AS / S /8 6 3/ / A-AS /8 6 5/ A-AS / S /8 10 3/ / A-AS / S /4 10 3/ / A-AS / A-AS / A-AS / A-AS / S / A-AS / / A-AS / A-AS / A-AS / A-AS / / A-AS / A-AS / S / / A-AS / / A-AS / / A-AS / A-AS / S /8 24 3/ / Overall Length Tons R22 (+40oF) Holding capacity (LBS)4 Model Fitting Overall Length Tons R22 (+40oF) Holding capacity (LBS)2 Notes: 1) R22 at 40 o F evaporator. 2) R22. 3) R22 at 40 o F divided by ) R22 at 0 o F saturation. Table D.24: Suction accumulator cross reference chart D - 19

91 Figure D.3 : R-22 refrigerant velocity chart D - 20

92 Figure D.4 : R-22 refrigerant pressure drop chart D - 21

93 R22 Thermophysical Properties. Properties of Saturated Liquid and Saturated Vapor. Temp. Pres. (MPa) Density (kg/m 3 ) Volume (m 3 /kg) Enthalpy (kj/kg) Entropy (kj/kgk) Spec. Heat, c p (kj/kgk) ( o C) ( o C) Liquid Vapor Liquid Vapor Liquid Vapor Liquid Vapor Vapor Liquid Vapor Liquid Vapor Liquid Vapor b c p /c v Sound velocity (m/s) Viscosity (mpas) Thermal Cond. (mw/mk) Surface tension (mn/m) Temp. Note : b = Normal boiling point. c = critical point. Table D.25: R-22 Thermophysical properties D - 22

94 R22 Thermophysical Properties. Properties of Saturated Liquid and Saturated Vapor. Temp. Pres. (MPa) Density (kg/m 3 ) Volume (m 3 /kg) Enthalpy (kj/kg) Entropy (kj/kgk) Spec. Heat, c p (kj/kgk) ( o C) ( o C) Liquid Vapor Liquid Vapor Liquid Vapor Liquid Vapor Vapor Liquid Vapor Liquid Vapor Liquid Vapor c p /c v Sound velocity (m/s) Viscosity (mpas) Thermal Cond. (mw/mk) Surface tension (mn/m) Temp. Note : b = Normal boiling point. c = critical point. Table D.26: R-22 Thermophysical properties D - 23

95 R22 Thermophysical Properties. Properties of Saturated Liquid and Saturated Vapor. Temp. Pres. (MPa) Density (kg/m 3 ) Volume (m 3 /kg) Enthalpy (kj/kg) Entropy (kj/kgk) Spec. Heat, c p (kj/kgk) ( o C) ( o C) Liquid Vapor Liquid Vapor Liquid Vapor Liquid Vapor Vapor Liquid Vapor Liquid Vapor Liquid Vapor c p /c v Sound velocity (m/s) Viscosity (mpas) Thermal Cond. (mw/mk) Surface tension (mn/m) Temp. Note : b = Normal boiling point. c = critical point. Table D.27: R-22 Thermophysical properties D - 24

96 R22 Thermophysical Properties. Properties of Saturated Liquid and Saturated Vapor. Temp. Pres. (MPa) Density (kg/m 3 ) Volume (m 3 /kg) Enthalpy (kj/kg) Entropy (kj/kgk) Spec. Heat, c p (kj/kgk) ( o C) ( o C) Liquid Vapor Liquid Vapor Liquid Vapor Liquid Vapor Vapor Liquid Vapor Liquid Vapor Liquid Vapor c p /c v Sound velocity (m/s) Viscosity (mpas) Thermal Cond. (mw/mk) Surface tension (mn/m) Temp. Note : b = Normal boiling point. c = critical point. Table D.28: R-22 Thermophysical properties D - 25

97 R22 Thermophysical Properties. Properties of Saturated Liquid and Saturated Vapor. Temp. Pres. (MPa) Density (kg/m 3 ) Volume (m 3 /kg) Enthalpy (kj/kg) Entropy (kj/kgk) Spec. Heat, c p (kj/kgk) ( o C) ( o C) Liquid Vapor Liquid Vapor Liquid Vapor Liquid Vapor Vapor Liquid Vapor Liquid Vapor Liquid Vapor NIL NIL c c p /c v Sound velocity (m/s) Viscosity (mpas) Thermal Cond. (mw/mk) Surface tension (mn/m) Temp. Note : b = Normal boiling point. c = critical point. Table D.29: R-22 Thermophysical properties D - 26

98 Figure D.5 : R-22 p-h diagram D - 27

99 Suva 407C R E FR IGE R ANT VELOCITY IN LINES (65 F Evap. Outlet) Figure D.6 : R-407C refrigerant velocity chart D - 28

100 Suva 407C R E FR IGE R ANT PRESSURE DROP IN LINES (65 F Evap. Outlet) Figure D.7 : R-407C refrigerant pressure drop chart D - 29

101 R407C Thermophysical Properties. Properties of Saturated Liquid and Saturated Vapor. Absolute Pressure (MPa) Temperature ( o C) Density (kg/m 3 ) Volume (m 3 /kg) Enthalpy (kj/kg) Entropy (kj/kgk) Spec. Heat, c p (kj/kgk) Bubble Dew Liquid Vapor Liquid Vapor Liquid Vapor Liquid Vapor Vapor Liquid Vapor b c p /c v Sound velocity (m/s) Absolute Pressure (MPa) Note : b = one standard atmosphere. c = critical point. Table D.30: R-407C Thermophysical properties D - 30

102 R407C Thermophysical Properties. Properties of Saturated Liquid and Saturated Vapor. Absolute Pressure (MPa) Temperature ( o C) Density (kg/m 3 ) Volume (m 3 /kg) Enthalpy (kj/kg) Entropy (kj/kgk) Spec. Heat, c p (kj/kgk) Bubble Dew Liquid Vapor Liquid Vapor Liquid Vapor Liquid Vapor Vapor Liquid Vapor c p /c v Sound velocity (m/s) Absolute Pressure (MPa) Note : b = one standard atmosphere. c = critical point. Table D.31: R-407C Thermophysical properties D - 31

103 R407C Thermophysical Properties. Properties of Saturated Liquid and Saturated Vapor. Absolute Pressure (MPa) Temperature ( o C) Density (kg/m 3 ) Volume (m 3 /kg) Enthalpy (kj/kg) Entropy (kj/kgk) Spec. Heat, c p (kj/kgk) Bubble Dew Liquid Vapor Liquid Vapor Liquid Vapor Liquid Vapor Vapor Liquid Vapor c p /c v Sound velocity (m/s) Absolute Pressure (MPa) Note : b = one standard atmosphere. c = critical point. Table D.32: R-407C Thermophysical properties D - 32

104 R407C Thermophysical Properties. Properties of Saturated Liquid and Saturated Vapor. Absolute Pressure (MPa) Temperature ( o C) Density (kg/m 3 ) Volume (m 3 /kg) Enthalpy (kj/kg) Entropy (kj/kgk) Spec. Heat, c p (kj/kgk) Bubble Dew Liquid Vapor Liquid Vapor Liquid Vapor Liquid Vapor Vapor Liquid Vapor c c p /c v Sound velocity (m/s) Absolute Pressure (MPa) Note : b = one standard atmosphere. c = critical point. Table D.33: R-407C Thermophysical properties D - 33

105 Figure D.8 : R407C p-h diagram D - 34

106 R410A Thermophysical Properties. Properties of Saturated Liquid and Saturated Vapor. Absolute Pressure (MPa) Temperature ( o C) Density (kg/m 3 ) Volume (m 3 /kg) Enthalpy (kj/kg) Entropy (kj/kgk) Spec. Heat, c p (kj/kgk) Bubble Dew Liquid Vapor Liquid Vapor Liquid Vapor Liquid Vapor Vapor Liquid Vapor b c p /c v Sound velocity (m/s) Absolute Pressure (MPa) Note : b = one standard atmosphere. c = critical point. Table D.34: R-410A Thermophysical properties D - 35

107 R410A Thermophysical Properties. Properties of Saturated Liquid and Saturated Vapor. Absolute Pressure (MPa) Temperature ( o C) Density (kg/m 3 ) Volume (m 3 /kg) Enthalpy (kj/kg) Entropy (kj/kgk) Spec. Heat, c p (kj/kgk) Bubble Dew Liquid Vapor Liquid Vapor Liquid Vapor Liquid Vapor Vapor Liquid Vapor c p /c v Sound velocity (m/s) Absolute Pressure (MPa) Note : b = one standard atmosphere. c = critical point. Table D.35: R-410A Thermophysical properties D - 36

108 R410A Thermophysical Properties. Properties of Saturated Liquid and Saturated Vapor. Absolute Pressure (MPa) Temperature ( o C) Density (kg/m 3 ) Volume (m 3 /kg) Enthalpy (kj/kg) Entropy (kj/kgk) Spec. Heat, c p (kj/kgk) Bubble Dew Liquid Vapor Liquid Vapor Liquid Vapor Liquid Vapor Vapor Liquid Vapor c p /c v Sound velocity (m/s) Absolute Pressure (MPa) Note : b = one standard atmosphere. c = critical point. Table D.36: R-410A Thermophysical properties D - 37

109 R410A Thermophysical Properties. Properties of Saturated Liquid and Saturated Vapor. Absolute Pressure (MPa) Temperature ( o C) Density (kg/m 3 ) Volume (m 3 /kg) Enthalpy (kj/kg) Entropy (kj/kgk) Spec. Heat, c p (kj/kgk) Bubble Dew Liquid Vapor Liquid Vapor Liquid Vapor Liquid Vapor Vapor Liquid Vapor c c p /c v Sound velocity (m/s) Absolute Pressure (MPa) Note : b = one standard atmosphere. c = critical point. Table D.37: R-410A Thermophysical properties D - 38

110 Figure D.9: R410A p-h diagram D - 39