Modern centrifugal compressors using double-end, variable-speed electric motors are high performance and reliable options for pipeline compression stations. Such a compressor train is equipped with two compressor casings coupled through two gear units on either side of a double-end variable speed drive (VSD) motor.
Alternatively, a high-speed VSD electric-motor can be used without any gear unit or just one gear unit for the high-pressure casing. A unit of this type can easily compress a large volume of gas from relatively low pressure to pressure above 100 Barg (for modern high-pressure pipelines, which can transfer gas to long distances), using a single train of two compressor bodies (casings), each casing has eight or nine impellers.
Modern electric, motor-driven compressors could offer high reliability and availability if designed, manufactured and operated properly. As an indication, a single shutdown every five years would not be unusual.
A single-end electric motor coupled to two or three compressor casings was the traditional design for multi-casing centrifugal compressors. One of the problems with single-end electric motors is the first compressor casing has become the governing unit, which has resulted in many operational problems. Maintenance and access to the casings at the middle of the train have caused many difficulties and not been recommended for modern multi-casing compressors.
VSD Double-Casings Vs. Integrally-Geared
For gas pipeline compression applications, which require dry gas seals, a VSD double-casings compressor design might be a better option than integrally geared compressors, since a comparable compressor might require eight or 10 separate dry-gas seals, compared to four in a double-casings design.
In addition, variable-speed operation that could be in 70%-105% range (of nominal speed) for a double-casings machine using a double-end VSD electric motor might offer greater operational flexibility.
The compressor string with two gearboxes, two compressor casings and a VSD electric motor might result in a complicated torsional system. In cases in which torsional vibration frequencies coincide with resonance frequencies, large torsional or internal stresses could be generated. However, overall the dynamic and torsional situation is less complex compared to integrally geared machines. Continuous operations under resonance conditions could result in fatigue failure. Free-vibration analysis and forced-response studies should be conducted for an analytical review of rotor response to static and harmonic torsional loads.
In alternative configurations, high-speed VSD electric-motors with one gear unit or direct-drive arrangement, the dynamic situation is usually more favorable. The elimination of one or both gear units can ease the situation; such a design can offer better operation, efficiency and reliability.
VSD electric motors generate pulsating torques; even if the pulsating torques are extremely small with respect to the main torque, they can excite compressor train resonances with potential shaft or couplings damages.
The VSD double-casings compressor can be designed to improve access to bearings, seals and rotating assemblies. This compressor arrangement does not require the station piping to be removed for normal maintenance access. Also nozzle loads can be two or three times the American Petroleum Institute (API)-617 loads.
All these can offer better piping, access and maintenance compared to integrally geared machines that mandate much lower nozzles loads (API-617 loads or lower). My estimate is nearly 60% of large integrally geared compressors in critical services have one or more nozzle with expansion joints.
Integrally geared compressors lend themselves to use in many interstage coolers, in particular water-cooled, heat-exchangers units, combined in a compact compressor arrangement. In other words, for a typical integrally geared compressor, the coolers should be supplied by the vendor. In such a compressor package, coolers are often marginally designed (with minimum margins), and the interstage piping and associated pressure drops can be reduced by locating the intercoolers close to the compressor.
For many pipeline gas station projects (nearly all of them), the use of water cooling is not an option, so air-coolers should be used. In some other pipeline projects, the number of cooled sections (the number of coolers) should be minimized. For high-pressure gas pipeline projects, a double-casings two (or three) cooled sections might be a better option compared to a four (or five) cooled-sections integrally-geared compressor package.
In some gas pipeline station projects, coolers should be designed and fabricated by specialized cooler manufacturers with large margins and sophisticated designs. In other words, there is a necessity to exclude coolers from the compressor vendor scope – the coolers should be ordered separately to a specialized cooler vendor under the direct supervision of the purchaser.
Obviously in all the above-mentioned options, coolers manufactured by independent vendors cannot be properly coordinated for inclusion in a compact compressor package of a typical integrally geared machine design. For such pipeline projects this type of compressor could not be an option.
Usually VSD double-casings compressors could offer better operation, flexibility, reliability and arrangement for high-pressure gas pipeline applications and are recommended for many pipeline compression services rather than integrally geared compressors.
The best way to select between these two options is through evaluation of cost of purchase and operation. In my experience, almost 90% of all high-pressure gas pipeline applications, a VSD double-casings compressor is the better choice.
Gas Pipeline Considerations
For many pipeline services, the risk of condensation downstream of the compressor units is a concern. The liquid can result in corrosion and erosion. To mitigate this risk, the items and substances that could condensate need to be removed from the gas stream. To maximize the removal of such substances, the compressor’s intermediate stage pressure might need to operate within a fixed pressure range. This type of a gas pipeline requirement can impose restrictions on the compressor arrangement and design.
Substances, such as hydrogen sulfide (H2S) in the CO2, could cause problems in some gas pipeline services. The presence of such can increases the complexity of the compressor material and seal selections. The material selection is critical for such gas services. For instance, the presence of H2S can lead to stress corrosion cracking (SCC).
Another risk is the rapid depressurization in the high-pressure gas pipeline services. For many gas services above 130 Barg operate near normal ambient temperatures. Rapid depressurization can result in temperatures as low as -60° C. The most important thermodynamic properties to predict centrifugal compressor performance are density, speed of sound and specific-heat (“Cp” and “Cv”).
VSD double-casings compressors for high-pressure applications are special machines. Sometimes a vendor may specify vibration limits more than the API-617 for these machines. However, the best recommendation is to follow the API-617 limits. Usually values close to API limits would be acceptable. For some high-pressure special compressors vibration values, 5% or 10% more than API limits might also be acceptable after careful evaluations, particularly if it is within previously successful experience ranges. However, vibration limits of more than 1.1 times the API limits should not be accepted.
Dry Seal Gas Systems
In most cases, an external seal gas source is not available at the conditions required for the VSD double-casting compressor applications. For such compressors a dedicated seal gas booster (small compressor) system should be used. This gas-booster system is comprised of a seal-less reciprocating compressor or seal-less centrifugal vertical compressor that does not allow the sealing gas to make contact with the external environment.
During standstill conditions, the use of seal gas from the compressor discharge is not possible, and the dry-gas-seal is at risk of contamination, especially if the gas inside the machine is dirty, wet or contaminated (by condensations, dirt, etc.). Normally to mitigate this risk, the compressor would be depressurized after a trip. The installation of the seal gas booster allows the compressor to be kept pressurized after a trip, while the dry-gas seals remains correctly buffered and ready to restart. In such situations, the seal system heaters should also be working to prevent liquid formations in the seal gas. This design avoids the need to depressurize the compressor during trips. All the seal gas support system components should usually be modularized on two separate seal skids (seal panels), for each casing.
Final validation of a double-casings VSD compressor train should be done at its manufacturing shop with a full-speed, full-load test and American Society of Mechanical Engineers (ASME) PTC-10 Type-1 performance test. During this testing, torque pulsation measurements should be conducted for the validation of torsional analytical results and pulsation amplitudes under steady state and transient conditions (such as start-up).
For the performance test, both the head and efficiency curves should be respected and evaluated. Too often a test performance curve, both for head and efficiency, is slightly lower than the theoretical expected curve. This is because the current sophisticated theoretical and simulation formulations used by experienced vendors are a way to simulate the compressor performance accurately, but slightly higher than the actual performance.
Therefore, the performance test should result in a slightly lower than the expected curves. However, there have been exceptions, and the test result might be slightly more than expected. If there is a significant mismatch, there could be a serious mistake in design.
If all stages show slightly lower or higher head compared to expected head, a speed adjustment may be required. This is a great advantage of a VSD driven compressor package. In a case study, the rated speed of compressor was 1.5% reduced to achieve the required performance curve, the head required and the optimum efficiency.
Sometimes one or more stages of a compressor could show extra-head and polytrophic efficiency. Such a situation is usually more difficult to deal since whole the compressor train is operated at the same speed. Options such as trimming some impellers have been proposed with great promises; these modifications are expensive, risky options, best left for special cases.
Often different stages balance each other and problems could only be at operating cases, far from rated cases. A solution for such a situation might be limiting the overall map of the double-casings compressor. Practically speaking, this might not be a real concern for many compressor packages. Regarding the left side of a compressor performance curve (high-pressure section), the final discharge pressure is usually limited by a high-pressure trip and on a higher pressure set point by pressure relief valves. On the left side of a compressor performance curve (over-flow section), there are always some limits on the maximum flow that the pipeline system can handle.
With respect to these cases, the following guidelines can be drawn for the bidding stage to order a compressor train that can deal with such a situation at the shop performance test:
• The speed variation range should be relatively wide. For example, 65-105% of normal speed (if possible) to deal with uncertainties in the performance test and any mismatch between the test (actual) performance and expected theoretical performance.
• The stage curves and overall performance curve should be properly wide. The overall performance curve at the bidding stage is just an expected one. This might be trimmed by the actual performance of the machine. Therefore, in theory it should be wide enough that in a worst-case scenario it will be restricted; it could still offer a reasonable operational flexibility.
Author: Amin Almasi is a lead mechanical engineer, specializing in rotating equipment and pipelines in Australia. He is a chartered professional engineer of Engineers Australia (MIEAust CPEng – Mechanical) and IMechE (CEng MIMechE) and holds a master’s and bachelor’s degree in mechanical engineering.