Reciprocating compressors receive and deliver
gas in discrete slugs (or pulses) as the pistons move back and forth. The
periodic piston movement causes pressure waves to propagate throughout the
compressor piping system. These pressure waves have a fundamental frequency
component that matches the compressor rotational frequency and its multiples
(also called compressor harmonics). The compressor speed is typically given in revolutions
per minute (rpm) and pressure frequencies are expressed in cycles per second or
Hertz (Hz)—compressor
speed is typically given in rpm and can be converted to Hz by dividing the rpm
value by 60.
For a double-acting compressor, there will be
two pulses of gas for each revolution, resulting in higher excitations of the
second running speed order. As these pulses of gas enter or exit the compressor
cylinder, they will give rise to pressure pulsations that move in all
directions at the speed of sound down the pipe. These pulses can be quite large,
so attenuation devices generally called “pulsation suppression dampers” or
“dampeners” are typically mounted on the inlet and discharge of the compression
cylinders, as shown in FIG.
1.
These pressure waves are most damaging if one
or more piping acoustic resonances are close to a compressor piston excitation
frequency. These high-pulsation levels can lead to high unbalanced shaking
forces when they act on pipe bends, elbows, branches and nearby vessels. These
shaking forces are dynamic in nature and can be most damaging if the shaking
forces frequency coincides with the piping mechanical resonance frequencies,
which in turn can lead to high cycle stress fatigue and failure. Furthermore,
the uncontrolled pulsations can cause a decrease in compressor efficiency due
to an increase in power consumption or a reduction of capacity. Pulsations can
also result in errors in flowmeter readings and process control systems and
increase site noise levels. The primary objective of a pulsation study during
the design phase of a capital project is to minimize the effect of uncontrolled
pulsation waves on the downstream vessels and connected compressor piping.
American
Petroleum Institute (API) 618 design approaches for the pulsation study. There are three design
assessment levels covered in API 618. These three approaches are detailed in
this section and include:
Design Approach
1: Design Approach 1 involves
empirical pulsation suppression device sizing and the use of proprietary and
empirical analytical techniques to meet line-side pulsation levels and the
maximum pressure drop based on compressor normal operating conditions. The main
objective of this approach is to evaluate the proper size and volume of
pulsation suppression vessels in the inception of the proposal stage of a project
by the vendor before a detailed process design is performed.
Design Approach
2: Acoustic simulation of pulsation waves in piping system without
mechanical response analysis. For Design Approach 2, an acoustic simulation
is required to evaluate and control pulsation in piping systems for
reciprocating compressors. This pulsation simulation report contains a summary
of pressure pulsations in the piping system as well as modification proposals
to reduce pulsation to the agreed level as per API criteria. The original
system layout will be checked with respect to pulsation. All operating cases
that are characterized by steady-state operating conditions will be checked, therefore
transient cases (e.g., startup cases) are excluded. Different gas properties
and operation cases are considered. Valve unloading cases are considered as
operation cases. Measures are proposed to reduce pressure pulsation, such as
the installation of orifices, and changes in piping length or diameter. Shaking
forces are used in the subsequent vibration study. Normally, if the selected
dampers result in pulsation levels that exceed the API criteria, then proper
orifice plates (with suitable bores) should be selected by the vendor to solve
the problem and meet API design criteria.
Design Approach
3: Acoustic simulation of pulsation waves in piping system with
mechanical response analysis. Design Approach 3 requires that a mechanical
analysis with forced mechanical response analysis be performed in addition to the
acoustic study required in Design Approach 2. The forced mechanical response
analysis for the compressor cylinder as well as associated piping systems must be
performed using pressure pulsation results from the acoustic study.
Pulsation
frequency plots according to API guidelines. Pressure pulsation
plots represent the response of the system to the compressor’s moving piston
action. The plots are actually a series of plots corresponding to each
compressor harmonic for the rpm range used in the analysis. The harmonic number
is displayed near the top line of each figure to show the relevant compressor
harmonic. The harmonic plots tend to be wider for higher harmonics, as the
width is also a function of the harmonic number. For example, if the rpm range
varies from 900 rpm–990
rpm, the first harmonic will vary from 15 Hz–16.5 Hz and the second will vary from 30 Hz–33 Hz. The second
harmonic has a 3-Hz variation vs. a 1.5-Hz variation for the first harmonic.
Double-acting cylinders tend to have a higher even (2, 4, 6...) harmonic response,
while single-acting cylinders will have higher response at odd (1, 3, 5...)
harmonics. These pulsation plots give the piping designer additional
information for support design and spacing. The API 618 pulsation level is also
indicated on this diagram. API 618 sets two different guidelines: one is for compressor
cylinder flange pulsation up to the pulsation damper and the other is for piping
pulsations. FIG. 2
shows a typical pressure pulsation plot in accordance with API criteria.
API 618 cylinder
flange criteria.
Based
on API guidelines, unless other criteria (such as loss of compressor
efficiency) are specified, the unfiltered peak-to-peak pulsation level at the
compressor cylinder flange—as
a percentage of average absolute line pressure—shall be limited to < 7% or the value computed
from Eq. 6 in Sec. 7.9.4.2.5.2.1.1 This is intended to limit the
pulsation pressure at either the suction or discharge cylinder flanges of a
reciprocating compressor. Excessive pressures in this region promote improper
valve operation, leading to flutter or slamming against the seat and causing
higher compressor losses. The guideline for pressure pulsation between the cylinder
flange (orifice location) and pulsation bottles is defined in API 618. The
guideline equations are shown here in Eqs. 1 and 2:
where,
Pcf = the maximum allowable unfiltered
peak-to-peak pulsation level, as a percentage of average absolute line pressure
at the compressor cylinder flange
Pl = the average
absolute line pressure, bar
R = the stage
pressure ratio.
API 618 piping
pulsation criteria. This
guideline is used for general piping starting after pulsation bottles away from
reciprocating compressors. For systems operating at absolute line pressures
between 3.5 bar and 350 bar (50 psia–5,000 psia), the peak-to-peak pulsation level of each individual
pulsation component should be limited to that calculated by Equation 8 in Sec.
7.9.4.2.5.2.2.2.1 API 618 guidelines are used to determine an
allowable level of pulsation pressure (Eq. 2).
where,
P1 = the maximum
allowable peak-to-peak level of individual pulsation components corresponding
to the fundamental and harmonic frequencies, expressed as a percentage of mean
absolute line pressure
A = the speed of
sound for the gas, m/sec
PL = the mean absolute
line pressure, bar
DI = the inside diameter
of line pipe, mm
f = the pulsation
frequency, Hz.
What are
the unbalanced shaking forces? Pulsation pressures are normally a small
percentage of the piping’s static design pressure. The hoop stresses associated
with pulsations are normally insignificant with respect to the design stresses.
Problems arise when acoustic pulsation acts directly on a mechanical system and
results in a shaking force large enough to cause vibration. For example, a high-pulsation
pressure in the middle of a pipe run will only excite piping shell resonance
and will normally not cause a significant lateral vibration problem; however, a
high-pressure pulsation at an elbow can result in a dynamic shaking force.
Typically, shaking forces arise when a dynamic
pressure is present at a change in pipe direction or area.2 System shaking
forces are represented as pulsations acting on a set of nodes with area factors
related to the respective node’s geometry. When nodal pressures are multiplied by
the effective areas, the nodal shaking forces are determined. The general shaking
force is defined by the following relationship (Eq. 3):
In the above equation, k is the shaking
force number and N
is the number of piping nodes. Since vectorial forms of pressures and forces
are used here, node force components can have both positive and negative values,
depending on their phase orientation of pressures and forces over time. By
design, there will be cases when negative and positive values can completely cancel
out to minimize shaking forces. Also, note that at 45° piping bends, 70.7% of the dynamic gas pressure
force will act in the direction of flow and 70.7% will act normal to the
direction of flow.
To calculate the maximum vibration and cyclic
stress levels in a pipe, the maximum pulsation levels are taken from the
acoustical analysis and plugged into the mechanical response model. The stress
intensification factors calculation is performed according to American Society
of Mechanical Engineers (ASME) B31.3 to determine maximum piping stress levels.
If all vibration and stress levels are satisfactory, no further analysis is
required. However, if problems are found, the analyst must modify the piping
design until a viable system design is found.
How orifice plates can control pulsation on
dampeners to below API limitations. During the conceptual and design phase of a project, designers analyze their
compressor/piping systems to predict pressure pulsations levels under all
expected operating conditions. If issues are found, designers can then evaluate
the effects of the various piping lengths, orifices and pulsation dampers on pulsations
levels in an effort to stay under API limits at all times. However, once the compressor/piping
system elements are fabricated and installed, most methods of dealing with high
pulsations and excessive piping vibrations tend to be very costly or
impractical. For example, field changes to piping lengths, replacing or
modifying pulsation dampeners, or adding additional piping restraints (FIG. 3) are likely to
be expensive and are considered to be “last straw” options.
For these
reasons, if field problems are experienced, the first attempt at reducing
pressure pulsations in the field usually involves strategically placing
orifices between existing flange sets. Fixed orifice plates are one of the most
common elements employed to address field pulsation issues because they are inexpensive,
relatively easy to install and effective in breaking down pressure pulses.
It is common to find orifice plates at multiple locations throughout the piping
system. Fixed orifices are thin metal sheets with a round hole of a specified
diameter located at the center of the pipe cross‐section. The orifice
plate is retained between two adjacent pipe flanges that are held together with
multiple threaded fasteners and sealed with flat gaskets. Once the flanges are
installed, the orifice plates remain in place and can only be removed or
changed by stopping the compressor, completely venting all gas to atmospheric
pressure, loosening all the flange-threaded fasteners, removing the original
orifice plates, installing new orifice plates with new gaskets, reassembling
and tightening the threaded fasteners, pressurizing the system with gas, and
restarting the compressor. FIG. 4 shows orifice plate locations and how they are
installed in piping and at the inlet spool of cylinders (photo on the right).
Takeaways. Reciprocating
compressors generate significant mechanical excitation and pulsation energy at
discrete frequencies that are transmitted throughout the machinery, vessels and
piping system. These excitations can occur at many different frequencies and
directions simultaneously and can change based on operating conditions. To
reduce the risk of uncontrolled pulsations and their associated destructive vibrations,
the mechanical characteristics of a compression system should be analyzed at
the design stage or when troubleshooting existing field problems.
For the mechanical analysis of a reciprocating
compressor manifold or skid, the system is typically modeled in detail using
finite element modeling techniques. In addition, a modal analysis should be
performed to predict the natural frequencies of the compressor system and the expected
mechanical resonance responses of system components. A forced response analysis
will be performed to estimate peak stress and vibration levels to ensure
acceptability. Piping and restraint system modifications should be made during
the design phase to avoid any coincidences between the natural frequencies of
the piping and the frequencies of residual pulsation energy. This analysis is
discussed in more detail in the API 618 and API 688 standards. HP
LITERATURE CITED
MOHSEN
ESHAGHI
leads the rotary department of a company that manufactures API-compliant
compressor packages. He has 8 yr of experience in research, design and
construction in oil and gas industry projects. He is a lecturer of machinery
courses and has authored articles about field operations and maintenance of
API-compliant compressors. Eshaghi earned B.Sc. and M.Sc. degrees in mechanical
engineering. The author can be reached at mohsen.eshaghi.job@gmail.com.
ROBERT PEREZ is a mechanical engineer with more than 40 yr of rotating equipment experience in the petrochemical industry. He has worked in petroleum refineries, chemical facilities and gas processing plants. He earned a BS degree in mechanical engineering from Texas A&M University at College Station, Texas, an MS degree in mechanical engineering from the University of Texas at Austin and holds a Texas PE license. Perez has authored numerous technical articles for magazines and conferences proceedings and has authored five books and coauthored four books covering machinery reliability. He is also the Technical Editor of Kane’s Rotating Machinery Dictionary.