Refrigeration Plant Noise Control
Most operators are aware that noise control may be necessary with respect to their engine driven compressor plants in order to comply with the AEUB Noise Control Directive ID 94-4. However noise control of refrigeration plants raises some unique problems and some special applications.
In the case of a refrigeration plant where the compressors are driven by gas engines, then the main acoustic considerations apply as for a standard natural gas compressor with the addition that noise control may be necessary on the propane refrigerant condenser. Fan noise specifications are usually supplied by aerial cooler manufacturers with sound pressure levels referenced at 1 metre from the plane of rotation of each fan.
When dealing with fans, do not use the simple noise propagation equation found in the AEUB Noise Control Guidelines. The AEUB correctly state that this equation only applies to point noise sources. The fan is a plane source rather than a point source. Operators or engineers who use this equation to estimate noise at the nearest residences from the manufacturer's specified noise level at 1 metre from the fan, risk significant errors in their calculations - particularly if they are dealing with large diameter fans.
Manufacturers' sound pressure levels at 1 metre from the plane of fan rotation need first to be converted to overall sound power levels and appropriate directivity components applied. The engineer can then calculate the sound pressure levels at a given distance from the refrigeration plant. In the case of refrigeration plants with electrically driven compressors the problem of engine exhaust noise disappears. Typically the refrigerant condenser fan noise is the only source that will affect local residences.
If the operator is concerned about local noise protection for plant operators (i.e. OSHA concerns for noise levels in operating plant buildings rather than noise levels at residences some distance from the facility) then the following qualitative considerations should be noted.
- Screw & centrifugal refrigerant compressors are generally more noisy than reciprocating machines for the same horsepower.
- Perforated inner liners for the compressor building will be beneficial for all three types of machines but the benefit will be more noticeable in the case of screw and centrifugal machines due to the higher frequency noise spectra generated by these types of machines.
- Without perforated inner building skins typical noise levels 1 metre from an electrically driven compressor will be of the order:
- 95 to 100 dBA for reciprocating machines.
- 100 to 105 dBA for screw and centrifugal machines.
The noise generated by the screw machines is fairly sensitive to speed; the 3600 r.p.m. machines being noticeably noisier than the 1800 r.p.m. machines.
The perforated inner liner for the process building will have little effect on the 1 metre noise levels (i.e. immediately adjacent to the machines) but will typically reduce average noise levels in these process buildings by 2 - 3 dB for reciprocating machines and by 3 - 5 dB for the screw and centrifugal machines.
Optimize Refrigeration & Liquid Recovery in Winter
Many plant operators are concerned about their refrigeration systems in the summer. Hot weather causes an increase in required refrigerant condenser pressures with resulting reduction of refrigerant compressor performance. Increasing condenser pressure indirectly causes an increase in chiller pressure as the overworked refrigerant compressor cannot remove boiling refrigerant vapours as fast in hot weather conditions.
The simple refrigeration process is depicted on the pressure-enthalpy diagram below.
The refrigerant compressor takes the chiller vapours from point A to point B; the condenser from point B to point C; the flash valve from point C to point D; and the chiller revapourizes the refrigerant along line D to A. The horizontal projection of the line AB represents compressor horsepower and the horizontal projection of the line DA, represents the thermal capacity of the chiller. When chiller pressure rises, the boiling temperature also automatically rises with the result that plant dew point rises and plant liquid capacity decreases.
In the winter all the adverse parameters that affected summer performance are supposed to be reversed. Condenser pressure should drop from point B to point F with a result that the refrigerant compressors can now dedicate their horsepower to drawing down chiller pressure to point E. The complete process p-h loop in the winter is represented by EFGH.
The theoretical benefit experienced by the chiller in winter is two-fold.
- Firstly the chiller operating line HE is at a lower pressure and therefore at a lower temperature.
- Secondly the heat capacity of the chiller in winter (HE) is greater than its capacity in summer (DA).
Much of this theoretical winter benefit disappears when operators maintain refrigerant accumulator pressures unnecessarily high in winter. The accumulator pressure is often maintained by a hot refrigerant vapour by-pass around the condenser.
This pressure control is often misunderstood by operators and engineers. We offer below typical questions and answers pertaining to this pressure control point.
Q: Don't I need to maintain a high accumulator
pressure in order to maintain a high DP across the refrigerant flash
valve?
A: Some operators think that this high DP across the flash valve is
necessary for a high DT - therefore generating deeper levels of refrigeration.
In fact, we are concerned about the absolute downstream temperature,
not the temperature differential across this valve. By maintaining artificially
high accumulator pressure, refrigerant liquid starts the flash from
a higher temperature in the first place.
Q: When I try to lower the accumulator
pressure in winter, the accumulator floods with liquid. Why?
A: This is because, under vapour by-pass conditions, the condenser is
subcooling propane. Once the by-pass pressure maintenance is altered
downwards, this subcooled propane dumps from the condenser into the
accumulator. This is not a problem, it just means that the accumulator
surge level of propane will rise when operator reduces the accumulator
set pressure.
Q: I need a large DP across the flash
valve in order to provide enough hydraulic drive to maintain the level
in the chiller.
A: Change to a larger port on the flash valve.
Q: I have an interstage economizer in
my system. If I allow the accumulator pressure to drop won't this negate
the effect of the economizer?
A: To some extent this is true. For a two stage refrigerant, the
ideal economizer pressure is the geometric mean of the compressor suction
and compressor discharge pressure expressed in PSIA. For example if
a two stage refrigerant compressor is operating from 20 psia to 180
psia (for an overall ratio of 9:1) then the economizer should be set
to operate at 60 psia (dividing the two stages into equal ratios of
3:1). This means as operator drops his accumulator pressure he should
also reduce the economizer pressure.
Q: I need to maintain high accumulator
temperature and pressure so that my propane subcooler adequately preheats
my stabilizer feed.
A: Many refrig plants are designed this way for optimal summer
performance but it decreases the refrigeration capacity in winter. Consider
adding a feed/bottoms exchanger to the stabilizer train.
