Axial Compressors For Pipeline Compression Stations

January 2013, Vol. 240 No. 1

Amin Almasi, WorleyParsons Services Pty. Ltd.

This article highlights new developments and lessons learned from recent experiences in the pipeline industry in relation to 1) the selection, 2) the design, 3) the manufacturing, 4) the anti-surge systems and 4) the operation of axial compressors.

Axial-flow compressors are advanced turbo-compressors used for high capacities. They considerably reduce the weight, the size and the cost of the compressor package. They offer a very compact and modern compressor solution. The axial flow is highly sensitive to the profile shape/design of rotating and static components of an axial compressor. Another important factor is the rotor stability towards (near) stall/surge conditions, which is directly linked to the axial compressor operating range.

In recent years, the growth in axial compressor applications in pipeline projects has been driven by following main factors: 1) improvements in the 3-D blade/rotor designs; 2) modern materials and blade design/ manufacturing techniques; 3) a deep understanding of the aerodynamics of axial flows in axial-compressors; 4) advanced experimental methods/data analysis for the axial compressors (such as optical measurement techniques); 5) modern/complex anti-surge systems especially designed for axial compressors; and 6) new technologies for casing treatment, gas injection and bleeding.

There are complicated issues related to different design and operation aspects of axial compressors. Some examples could be: 1) the surge, 2) the stall, 3) the fragile blades, 4) the manufacturing problems, and 5) noise-related concerns. Figure 1 shows an example of an axial compressor.

Figure 2 shows an example of a turbo-compressor containing the axial compressor stages and the centrifugal compressor stages. The low pressure initial stages (high flow) are axial. The final two stages (the low flow stages) are centrifugal impellers.


Figure 2: An example of a turbo-compressor containing the axial compressor stages and the centrifugal compressor stages. The low pressure initial stages (high flow) are axial. The final two stages (low flow) are centrifugal impellers.

Axial Compressors For Gas Pipeline
In modern gas pipeline compression stations, the axial compressors are required to fulfill a relatively wide operating condition (an operating capacity/pressure range). Considering the steep nature of an axial compressor performance curve (the head vs. the flow), it is a great challenge. Sometimes, a gas pipeline operator requires an axial compressor to operate relatively far from its rated conditions. Modern methods such as the variable-speed and the variable “Inlet Guide Vanes” (IGV) systems should be used to provide an additional flexibility in the operation. Considering the required operational flexibility ( a. the speed variation, b. the variable IGV, c. others), many issues should be evaluated including structural, vibrational, weight, cost, manufacturability, accessibility, and reliability issues.

The operating Mach number is usually less than 0.8 for a subsonic cascade, but can go up to 2.0 and more at the tip of a transonic blade assembly. Some modern subsonic axial stages can develop pressure ratios in the order of 1.5-1.8. The transonic stages operate with pressure ratios around 2.0 and more while maintaining an acceptable efficiency and aerodynamic design. In a well-designed subsonic axial stage, the polytrophic efficiency could reach around 0.9. The polytrophic efficiency for a transonic blading assembly is a bit lower (say around 0.82-0.89).

High peripheral mean stage rotor velocities, around 300-340 m/s for subsonic rotors and up to about 580 m/s for transonic ones, could be achieved. An annulus radius ratio (hub-to-tip ratio) Rhub/Rtip, is usually chosen between around 0.45 (front stages) and approximately 0.9 (rear stages). The hub-to-tip ratios come from a careful optimization considering the aerodynamic, technical, mechanical and economic constraints. For inlet stages, the Rhub/Rtip values between 0.45-0.65 could be assigned, while outlet stages often are given a higher value of Rhub/Rtip (too often, from 0.75 to 0.9) in order to achieve a relatively high Mach number.

A proper axial compressor design should avoid a flow separation inside the machine. The axial Mach number distribution along the different axial blade stages should be considered carefully. The Mach number distribution should follow an acceptable pattern. The variation should not exceed a certain level.

The ultimate goal of an axial compressor design is to create an axial blade arrangement with the maximum pressure rise and the minimum total pressure loss (a relatively high efficiency) along with an acceptable operating range. The shapes of the different blades and component profiles play an important role because the profiles can affect the nature of the boundary layers and therefore the amount of losses (and the operating margins). The stage arrangement and the stacking are critical.

The maximizing of the adiabatic efficiency could have an important impact on the choice of stage geometrical and functional variables. On the other hand, the optimization of the surge/stall margins should be involved.

For an optimum axial compressor design (achieving an optimum weight and a compact compressor design), the number of stages should be optimized (usually the minimum possible stages, which can increase individual stage loading). This is also a factor that can affect the choice of the blade shape and the cascade parameters.

The availability of advanced materials for axial compressor blade/component construction (and high-quality production methods) makes it possible to reach levels of aerodynamic loading never experienced in traditional axial compressors, while preserving high levels of efficiencies for a normal and alternative operation cases. This is true both for the high-speed-subsonic design and the ultra-high-speed-transonic blades.


Axial Compressor Vs. Centrifugal Compressor

A centrifugal compressor offers a relatively wide operating range, a low maintenance and a high reliability. It is a very rugged and reliable machine.

An axial compressor offers a higher efficiency, a higher speed capability, and a higher capacity for a given size compared to a centrifugal compressor. However, an axial compressor has a relatively low pressure capability, a narrower operating range and fragile components.

The reality is the large, horizontally-split centrifugal compressors are manufactured for high capacity ranges. Some pipeline operators always prefer rugged, versatile and reliable centrifugal compressors to dedicated, efficient (but fragile) axial compressors. This led to very large size casings for low-pressure horizontally-split centrifugal compressors.

It is difficult to give a general rule for the selection to be made between a very large centrifugal compressor (with a rugged and massive design) and a compact, relatively-more-efficient, properly-optimized, more-economic axial compressor. In the gas industry:
1) Some operators prefer to use large centrifugal compressors (compared to the axial compressors) because of the rugged and more reliable designs.
2) For some very large capacities, the axial compressor could be the only available option.

Anti-Surge Systems At Axial Compressors

The delicate and fragile blades of axial compressors can easily be damaged (or even separated) as a result of a surge event. Also, axial compressors offer a complex surge line mapping. The surge line could change with a slightly different gas condition (or a slightly different gas composition). Axial compressors require dedicated anti-surge systems including five protection arrangements: 1) the anti-surge valves, 2) the hot-gas-bypass valves, 3) the inter-stage bleed valve (IBV), 4) the inlet guide vanes (IGV) system, and 5) the speed variation.

For axial compressors, the speed reduction and the IGV (inlet guide vane) characterization should be used to map the “surge-area” in a two-dimensional (2D) area. In addition to the anti-surge valves, the axial compressor is most often protected by an additional hot-gas-bypass (HGBP) recycle-loop. Usually a form of hot-gas-bypass valve is required for an axial compressor system. Particularly, it is mandatory in case that the anti-surge valve is not taken immediately after the compressor discharge. The axial compressor inter-stage bleed valve (IBV) should also be opened to provide sufficient flow to the compressor suction to avoid the surge at initial stage(s).

For an axial compressor, dynamic performance data such as the “Inlet Guide Vane” (IGV) stroke speed, the control and actuator delay, the valve stroke time, and other system dynamics should be involved in the dynamic model/simulation based on accurate and actual information. These data are critical and play important roles in axial compressor system dynamic simulation results, an anti-surge system design, reliability and overall safety. The dynamic model of an axial compressor should be accurately validated considering the criticality of the surge and disastrous consequences of a surge event in an axial compressor.

The IGV stroke time is usually in the range of four to eight seconds. Conceptually, fast responses of the IGV system might be considered desirable as they could help unload the axial compressor quickly. However, the IGV stroke could affect the compressor performance curve (for example, the distance between the operating point and the surge). Results from some dynamic simulations indicated the fast closing of the IGV mechanism (depending on the axial compressor operating map) could sometimes drive the axial compressor toward the surge.

Based on the simulation results, for some axial compressors, it is suggested that the IGV stroke time should be modified to an optimum value (instead of the fastest possible time value). Even for some axial compressors, a moderate IGV stroke time (say 8-12 seconds) may be better for the surge prevention. A stroke time of an inter-stage bleed valve (IBV) could possibly be as fast as two seconds.

However, for the axial compressors which the surge could initiate at final stages, an inter-stage bleed valve (IBV) fast opening could be problematic, since this action can significantly reduce the gas flow at the final stages. For the inter-stage bleed valve (IBV), an optimum window for the opening time appeared to exist to avoid the surge in either section of the axial stage (depending on the axial compressor design). All these optimum values should be identified using accurate dynamic simulations.

Stall At Axial-Compressors
As the axial compressor moves from the rated operating point to a near-stall operating point, the blade loading increases and the flow structures become stronger and unsteady. The tip leakage vortex can interact with the passage shock wave, leading to a self-sustained flow oscillation in the rotor passage. The results of near-stall operation could be: 1) a highly fluctuating fluid flow because of the shock/tip leakage vortex interaction, 3) a blade torque variation, and 3) a flow separation on the blade suction surface.

The abovementioned aerodynamic effects in a severe case event could result in a large blockage effect near the blading tip or even damage. There are two main types of rotating stall in axial compressors: 1) the long-length-scale (modal) stall and 2) the short-length-scale (spike) stall inception.

The modal stall inception is characterized by the relatively slow growth (over around 15-50 rotor revolutions) of a small disturbance of a long circumferential wavelength into a fully developed stall cell. The spike stall inception starts with the appearance of a large amplitude short-length-scale (two to three rotor blade passages) disturbance at the rotor tip (so-called “spike”), which could grow into a fully developed rotating stall cell within few rotor revolutions.

Casing Treatment, Gas Injection And Bleeding
Different patterns of hollow structures in the axial compressor casing are employed to improve the tip end-wall flow fields. These are referred to as the “casing treatments.” The interaction of the main flow with the flow circulating in these cavities (the hollow structures in the internal parts of the compressor casing) could have a positive impact on the compressor stability.

A proper casing treatment can widen the stable working range of an axial compressor and improve the efficiency. The near-stall conditions with a smooth casing (without a casing treatment) can induce vortex flows originating from the tip clearance flow crossing the tip gap. These vortex flows can hit the front part of the adjacent blade, indicating the possibility of a spill forward of low momentum fluid into the next passage. Also this flow feature could act as a trigger for the onset of rotating stall.

Using a proper casing treatment, the vortex trajectory remained aligned to the blade’s suction side. The disadvantages of the casing treatments are the space they need, the weight increase, a small efficiency penalty and the cost.

Using a proper casing treatment, an axial compressor can provide an increased stall-margin, with only small penalties of the efficiency. The well-known casing treatment methods are: 1) the circumferential groove-type treatment and 2) the slot-type treatment.

Regarding the groove-type casing treatment, generally a few (one, two or three) shallow grooves placed near the leading edge are better than numerous grooves all over the blade tip. Fewer and shallower grooves (if implemented carefully and correctly) can help to reduce the weight, the fabrication costs, and the loss generation. A single extended casing circumferential groove all over the blade tip section is also proposed as a successful design by some engineers. This can lead to the stability improvement since the flow relocation can help to relieve the locally severe blade loading.

In a typical slot-type casing treatment, the casing wall is circumferentially treated with a discrete number of axial rectangular slots over the blade tip section. In a successful casing treatment design, a significant increment in the flow stability compared to the solid casing wall is achieved. The stabilizing effects are based on the positive impact of casing treatment on the tip clearance flow and its resulting vortex.

A properly-arranged gas injection near the blade tips has proved to increase axial compressor stall margin, leading to a relatively wider operating range and a better operability. The tip injection increases the stability by unloading the rotor tip. This tip injection decreases the blade loading at the tip, allowing increased loading at lower blade spans before the blade stalls. A typical proposed design is to bleed pressurized gas from downstream of the rotor and properly inject it upstream. The tip injection could be an effective tool for recovering an axial compressor from a fully-developed stall. The bleeding (or aspiration) can be used to delay the blade flow separation.

Author:
Amin Almasi
is a lead rotating equipment engineer at WorleyParsons Services Pty Ltd, Brisbane, Australia and a frequent contributor to P&GJ. He previously worked at Technicas Reunidas (Madrid, Spain) and Fluor (various offices). He holds chartered professional engineer license from Engineers Australia (MIEAust CPEng – Mechanical), chartered engineer certificate from IMechE (CEng MIMechE), RPEQ (Registered Professional Engineer in Queensland) and he also holds MS and BS degrees in mechanical engineering. He can be reached amin_almassi @yahoo.com, amin.almasi @ymail. com.

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