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Lower condensing pressure 1
- Thread starter naifmbo
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The rating condition for high-pressure refrigerants (such as R-404A, -507 or -22) is a 40 degree evaporator temperature, 100-pounds delta P across the valve port, and 100 degrees liquid temperature. If the actual conditions are different than the nominal rating condition, then the actual valve capacity will be different than the nominal capacity.
The delta P across the TEV port is not simply the difference between liquid pressure and evaporator pressure. If there is a refrigerant distributor in the circuit, its delta P must be considered in the equation as well. The typical refrigerant distributor/tube assembly (in a high pressure refrigerant system), when correctly sized, will provide approximately a 35-pound delta P. The TEV delta P is calculated as follows: 226 pounds (liquid pressure) minus 45 pounds (the sum of 35 pounds distributor/tube assembly delta P plus 10 pounds evaporator inlet pressure), equals a 181-pound delta P across the TEV port. As the head pressure lowers, the available delta P across the TEV port is also lowered.
While lowering the head pressure results in a delta P reduction (which will decrease TEV capacity), this is accompanied by a lower liquid temperature (a result of the lower condensing temperature) which will increase the TEV capacity. The effect of lower delta P (reduced valve capacity) and lower liquid temperatures (increased valve capacity) will tend to negate each other without any significant change in TEV capacity.
While the lower head pressure yields reduced motor current and increased compressor efficiency, if lowered too far, eventually the TEV capacity would not be able to meet the demands of the evaporator load.
When this occurs, a portion of the evaporator will cease to transfer heat effectively, as liquid refrigerant would no longer be available to feed it. This will be evidenced by the higher superheat at its outlet. That portion of the evaporator that only sees refrigerant vapor essentially has become an extension of the suction line; it performs no useful work at all.
Reducing the TEV capacity, which leads to a starving evaporator, has, in effect, reduced the evaporator capacity. The end result is increased discharge air temperatures.
The delta P across the TEV port is not simply the difference between liquid pressure and evaporator pressure. If there is a refrigerant distributor in the circuit, its delta P must be considered in the equation as well. The typical refrigerant distributor/tube assembly (in a high pressure refrigerant system), when correctly sized, will provide approximately a 35-pound delta P. The TEV delta P is calculated as follows: 226 pounds (liquid pressure) minus 45 pounds (the sum of 35 pounds distributor/tube assembly delta P plus 10 pounds evaporator inlet pressure), equals a 181-pound delta P across the TEV port. As the head pressure lowers, the available delta P across the TEV port is also lowered.
While lowering the head pressure results in a delta P reduction (which will decrease TEV capacity), this is accompanied by a lower liquid temperature (a result of the lower condensing temperature) which will increase the TEV capacity. The effect of lower delta P (reduced valve capacity) and lower liquid temperatures (increased valve capacity) will tend to negate each other without any significant change in TEV capacity.
While the lower head pressure yields reduced motor current and increased compressor efficiency, if lowered too far, eventually the TEV capacity would not be able to meet the demands of the evaporator load.
When this occurs, a portion of the evaporator will cease to transfer heat effectively, as liquid refrigerant would no longer be available to feed it. This will be evidenced by the higher superheat at its outlet. That portion of the evaporator that only sees refrigerant vapor essentially has become an extension of the suction line; it performs no useful work at all.
Reducing the TEV capacity, which leads to a starving evaporator, has, in effect, reduced the evaporator capacity. The end result is increased discharge air temperatures.
I would add: maybe not as straightforward as this appears....
Most modern TXV's do not require much in terms of a pressure difference to support initial opening; but some older models do have a definite minimum inlet pressure.
Some arrangements also incorporate a fairly high differential check valve at the condensate line, which will get involved in estimating a minimum tolerable head pressure...
In terms of system efficiencies: Preserving ambient subcooling may be almost as beneficial as low discharge pressures...and mechanical subcooling may make better sense in terms of consistency of operation.
Fixed-Vi compressors don'y benefit as much from reduced discharge unless they are currently undercompressing. With such machines, subcooling is often a better solution than striving for low discharge pressures.
If this is a "compact" circuit, with similar evaporator duty and characteristics, then subcooling won't have a lot of benefit. With a more sprawling circuit, especially where substantial changes in elevation are involved, and pipe sizes could be marginal, low head pressure can decrease TXV and solenoid valve service life, because the liquid line pressure is lower but the volume gain due to liquid line flashing is greater, so the transition from a functioning and feeding TXV to one that is starving the evaporator tends to be very "sharp". Subcooling to a fixed temperature will take care of all that, though there is a greater investment involved in terms of insulated piping and heat exchangers.
Subcooling also increases the capacity of all the TXV's because the mass flow per unit heat moved at the evaporator is lower, and the pressure drop at the distributor will be lower....But the liquid distribution at the evaporator may become the next difficulty as the pressure drop at the distributor is required to balance the circuit.
In a larger plant: Experiment until you find the first load condition or target temperature to start suffering.
Most modern TXV's do not require much in terms of a pressure difference to support initial opening; but some older models do have a definite minimum inlet pressure.
Some arrangements also incorporate a fairly high differential check valve at the condensate line, which will get involved in estimating a minimum tolerable head pressure...
In terms of system efficiencies: Preserving ambient subcooling may be almost as beneficial as low discharge pressures...and mechanical subcooling may make better sense in terms of consistency of operation.
Fixed-Vi compressors don'y benefit as much from reduced discharge unless they are currently undercompressing. With such machines, subcooling is often a better solution than striving for low discharge pressures.
If this is a "compact" circuit, with similar evaporator duty and characteristics, then subcooling won't have a lot of benefit. With a more sprawling circuit, especially where substantial changes in elevation are involved, and pipe sizes could be marginal, low head pressure can decrease TXV and solenoid valve service life, because the liquid line pressure is lower but the volume gain due to liquid line flashing is greater, so the transition from a functioning and feeding TXV to one that is starving the evaporator tends to be very "sharp". Subcooling to a fixed temperature will take care of all that, though there is a greater investment involved in terms of insulated piping and heat exchangers.
Subcooling also increases the capacity of all the TXV's because the mass flow per unit heat moved at the evaporator is lower, and the pressure drop at the distributor will be lower....But the liquid distribution at the evaporator may become the next difficulty as the pressure drop at the distributor is required to balance the circuit.
In a larger plant: Experiment until you find the first load condition or target temperature to start suffering.
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