Optimizing Reciprocating Compressors for CPI Plants

Seems so long I don't updating this blog. So I decide before end of this years, I will profiding some useful information for all of you chemical engineering student about chemical process industry and optimization of the process. Hope you will enjoy this. our today discussion was about Reciprocating Compressors, you might think, what the heck is that ? I m suggest your read carefully this post.

Reciprocating compressors — the most commonly used type of compressor throughout the chemical process industries (CPI) — are flexible and efficient, and they can generate high head (from several bar to several thousand bar) independent of gas density. Worldwide, the installed reciprocating compressor horsepower is approximately two times that of centrifugal compressors.

However, the maintenance costs associated with reciprocating compressors are approximately three times greater than those for centrifugal compressors (due to valve, unloader and packing-maintenance requirements). This article provides practical recommendations for users to consider in an effort to improve the selection, operation and maintenance of reciprocating compressors in CPI applications.

Compressor designs

The Figure, shows the basic design of a reciprocating compressor. Close attention to the selection of the piston rod packing can improve performance, because this is a common source of reliability problems associated with reciprocating compressors, and is a common path for the leakage of potentially hazardous process gases. Experience shows that packing life can be extended by as much as a factor of three by adding the proper coating (tungsten carbide is a widely used coating material for piston rods).

Interstage cooling is required when the machine or gas being compressed has a temperature limit. In this case, as the gas cools, any liquid that may form is separated in interstage facilities and then the gas is returned to next compressor stage for further compression. Each compressor stage may consist one or more cylinders. Vendors usually offer a range of interstage pressures. The ability to optimize interstage pressures can help to minimize the total cost of ownership for the compressor and interstage facilities. This optimization can be done by evaluating the initial cost and operating costs of compressors and interstage facilities for various interstage pressures.

During operation, interstage pressures will increase during part-load operation (that is, operation at lower flow that results when an unloader device is used; this is discussed below) combined with variation in pressure at the suction inlet.

In a typical reciprocating-compressor design, the first stage may contain one or more cylinders and a clearance pocket. An additional bottle may be added to the cylinder with an actuated on/off valve. To avoid unwanted interstage pressure increases, users may consider installing additional clearance pocket(s) on the first-stage cylinder(s) and using part-load operation via the compressor control logic.

By selecting the right interstage design pressure, users can ensure proper operation in the face of part-load operation and variation in suction pressure. In general, the interstage design pressures should be around 15% higher than the interstage basic design values for applications that are working with common part-load steps (such as 25%, 50%, 75% and 100% capacity) and are expected to experience a variation in suction pressure of around +/- 7% during operation.

In some applications, reciprocating compressors must be designed to operate reliably in the face of considerable suction pressure variations while still providing full design flow at the desired discharge pressure. These operating requirements will have a direct impact on compressor sizing, especially the frame rating and motor power required for the unit.

The Graph shows load curves for the connecting rod of a reciprocating compressor in petroleum refining service. Variation in suction pressure (in this case, a roughly 7% reduction in suction pressure) results in a higher load on the rod. As a general rule, the compressor should be designed so that the maximum-anticipated rod load does not exceed 80% of the allowable rod load.

As shown on the y-axis in Figure, the rod load shall change sign from negative to positive and then negative again during one revolution of the crankshaft in order to provide proper lubrication for the mechanism (especially for the cross-head pin). The duration of rod sign reversal (the period during which load has the opposite sign) should not be less than 15 degrees of crank angle. The rod-load reversal peak (maximum amount of load in the reversed direction) should not be less than 3% of the actual combined load in the opposite direction. These minimum requirements should be satisfied under all possible operating conditions (especially in the face of suction-pressure variation and part-load operation, such as when an unloader device is used to decrease flow through the compressor).

In many cases, higher values of rod-reversal duration and peak are considered during compressor design to increase reliability. Minimum load-reversal duration corresponds to 50% capacity, and the reversal duration is more than 70 deg.

In general, the optimum speed for the reliable operation of reciprocating compressors is around 350 rpm. For compressors operating below 400 kW, speed on the order of 450 rpm is suitable. However, for those operating below 100 kW, higher speeds (even as high as 700 rpm) are acceptable.

Lubricated cylinders and packing may be preferred to extend service life. However, non-lubricated cylinder compressors should be used when the possibility of oil contamination cannot be tolerated in downstream operations (for instance, when trace amounts of lubricating oils could cause catalyst problems in downstream reactors).

For the optimum operation of reciprocating compressors, sufficient inertia — provided by a properly sized flywheel — is mandatory to regulate the variable reciprocating torque. On the graph bellow, shows brake torque versus crank angle for one revolution of the crankshaft for a reciprocating compressor used in a petroleum refinery setting. The red and blue curves represent compressor torque for normal full-load capacity and half-load (50%) capacity, respectively.

A step-less capacity-control system uses a hydraulically actuated, finger-type unloader. This device unloads the suction valve for only a portion of compression cycle to achieve the desired adjusted capacity.

A finger-type unloader has finger-shaped parts that act on the cylinder valve elements and keep them open for a defined duration during the compression cycle. Users should note that these finger-type unloaders have the potential to damage the valve-sealing elements and thus may have greater maintenance requirements.

A step-less capacity control system is recommended for larger machines (units rated above 2 MW, when large operation variation is expected). In these cases, step-less capacity control (working in the range of 20–100% capacity) is extensively used due to process requirements.

In general, valves and unloaders are responsible for nearly half (roughly 45%) of unscheduled reciprocating-compressor shutdowns, so valve and unloader selection can have a strong impact on reliability. And many consider the automatic cylinder valves to be the most critical components of such machines, as they are responsible for many unscheduled maintenance events. For large compressors (that is, those that operate at relatively low speeds with high pressure ratios), relatively large-bore ring-type valves (above 100 mm, or 4 in.) combined with plug-type unloaders should be considered first, to avoid reliability issues associated with finger-type unloaders. Since ring-type valves and plug unloaders are not available for smaller-sized compressors (those that operate at relatively higher speeds), such units typically use plate-type valves.

During operation, the rotating parts of the compressor, power transmission and driver will act like springs connected in series. This torsional dynamic system may create resonance (where one natural frequency of system coincides with one of excitation torque). In reciprocating compressor trains, there is always a risk of torsional resonance and torsional fatigue failure (that is, damage to component resulting from excessive cyclic loads).

Couplings that connect the driver to the compressor can be modified to tune the system to avoid torsional resonance. Several coupling options are available as follows:

1. Direct, forged-flange rigid connection (no coupling) between driver and compressor
2. High-torsional-stiffness coupling is allowed by torsional analysis. Since coupling options are limited, it may not be possible to find an acceptable coupling with the required torsional characteristics and service factor, especially for large compressors (above 3 MW)
3. Flexible coupling (which provides more elasticity and damping, but may require greater maintenance since elastic elements in such a coupling may need frequent replacement)
   

The most common reasons for problems caused by torsional vibration are lack of comprehensive torsional-vibration analysis, improper application and maintenance of couplings (especially flexible ones) and lack of appropriate monitoring. As a general rule-of-thumb, the shaft diameter of the electric motor should be equal to or greater than the reciprocating crankshaft diameter (because the crankcase is generally forged from a stronger steel grade compared to the motor rotor).

Condition monitoring

Condition monitoring, when done properly, can pay for itself by helping operators identify potential systems malfunctions at an early stage. A rigorous program should include monitoring of these important conditions:

Vibration (including continuous vibration monitoring of the compressor and motor casing, providing both alarm and shutdown capabilities):

• In general, velocity transducers are preferred over accelerometers (because interested frequencies for monitoring better match with velocity-measurement sensors). The optimum configuration for using a velocity transducer is to install one on each end of the crankcase, about halfway up from the base plate in line with a main bearing, both for compressor and motor
• Crosshead accelerometer (alarm)
  

Temperature:

• High gas-discharge temperature for each cylinder (with both alarm and shutdown capabilities)
• Pressure packing piston-rod temperature (alarm)
• High crosshead pin temperature (alarm), only for relatively large compressors (around or above 3 MW)
• High compressor main, and motor bearing, temperatures (alarm)
• Valve temperature (monitoring)
• Oil temperature, out of compressor frame (alarm)
• High jacket-water temperature of each cylinder (alarm)
  

In addition, proximity probes, typically located under the piston rods, provide alarm capabilities but are not used for shutdown. These are used to measure the rod position and determine wear or malfunctions. Such probes can quickly identify problems such as piston or rider band malfunctions, cracks in the piston rod attachment, a broken crosshead shoe or even a liquid carryover to a cylinder.
Improving maintenance

To support regular maintenance, the installation of any reciprocating compressor must ensure proper access to the entire compressor system, especially the non-drive end. In particular, adequate space and work areas must be provided to enable the complete withdrawal of the piston, removal of the cooler bundles or piping spool and laydown area (to carry out maintenance, dismantling of parts and repairs).

Similarly, three crane capacities must be properly identified: The total capacity of the overhead crane (to lift components for routine maintenance), the maximum maintenance weight (to ensure that the heaviest parts, usually the motor, can be lifted during overhauls), and the maximum installation weight (maximum skid weight, usually the compressor skid).

For a typical 7-MW API 618 compressor train for petroleum-refinery service, these crane capacities would be roughly 11 tons, 55 tons and 100 tons, respectively, and the required crane height would be roughly 12 m (around 40 ft)

Any time a given compressor must be stopped for an extended time, it should be turned a quarter-turn every week, using a barring device (this is a device that slowly turns the compressor to avoid locking and other problems that often arise during long stoppages of reciprocating compressors). A manual barring device can be used for relatively small compressors. A pneumatic barring device must be used for compressors rated above 750 kW (provided there is no area classification or power-availability problem).

For larger compressors (2 MW or larger), these special tools are often needed to carry out routine maintenance on reciprocating compressors. These tools cannot be easily purchased; they must be specially designed and fabricated based on the actual machine:

• Bearing extractor
• Piston extractor
• Valve extractor
• Piston fit-up tool
• Hydraulic tightening system
• Crosshead assembling tool
• Special lifting tools
• Partition plate-assembling tools
• Mandrels for wear bands
  

During maintenance of compressor mechanical components, the following criteria are important:

• Cylinder clearance for the outboard end should be around 4–6 mm (0.2–0.3 in.), and for the inboard end, clearance should be around 2–4 mm (0.1–0.2 in.)
• The allowable temperature of the machine bearings, piston rod, connecting rod bearing and crosshead should be maintained around 85ºC, and for the crosshead pin, it should be maintained around 90ºC.
• The vibration level of the crankcase should not exceed 100 microns, and the expected vibration level of the cylinders should be around 150 microns (these vibration recommendations are peak-to-peak vibration readings for an installed, trouble-free, middle-range machine around 1 MW).
• Bearings, piston rings and piston shoes should also be inspected regularly.
  

Auxiliaries and accessories

For auxiliaries and accessories, the optimum configuration is to install a local panel near the compressor skid (around 250 mm, or one foot away from the compressor skid), and on a standalone skid to minimize the potential for vibration damage.

The oil system should include two oil pumps, both sized for a capacity that is 20% greater than the maximum oil flow required for the compressor. At a minimum, two pumps should be used. Either a run-down tank (this is a stainless-steel tank that allows the supply oil to safely coast down the machine in the event that both pumps have failed), or a crank-shaft-driven main oil pump is required. Dual (two) removable bundle shell-and-tube oil coolers (TEMA C), double oil filters with a removable element and stainless-steel piping are also necessary.

Liquids should not be allowed to accumulate inside the compressor cylinder. For any application, a suction drum (sized appropriately with regard to application-specific retention time, flow velocity, and, if required, a mist-collection system to capture contaminants) with a drain provision should be provided. It may be part of pulsation control. To control pulsation, a vertical vessel is sometimes used as both suction drum and suction pulsation vessel, but this not recommended since vertical pulsation vessels cause relatively long piping to the compressor, which may lead to dynamic problems. Similarly, these are sometimes conflicting requirements for pulsation damping and suction separation, so this combined approach is not preferred. It may be used only for small compressors, let’s assume below 250 kW, with relatively light gases, such as those lighter than nitrogen.

Cooling-water systems are generally used as a heat sink for reciprocating compressors, to avoid hot spots and improve machine stability and reliability. To design a cooling system, first, the generated heat should be calculated. Then, the anticipated temperature rise should be identified. The cooling-water inlet temperature should be selected between 6°C and 16ºC above the inlet gas temperature. When selecting the pump to deliver cooling water to the compressor cylinder and packings, consider this:

• With regard to the slope of the operating curve (discussed in greater detail below), the selected operating point should not be in the flat or near-flat part of curve; rather, enough slope is needed for proper operation
• There should be a continuous rise from a selected operating point to shut off
• There should be proper shut-off pressure compared to a selected operating point (preferably 10% but a minimum of 6%)
  

Note: usually a larger size pump shall be selected to meet these recommendations.

Any reciprocating compressor system should be designed with a margin of excess flow capacity for the cooling system, to enable it to respond to situations that deviate from normal operation, where the need may arise later for additional cooling flow to remove excess generated heat (for example, unloaded operation when the compressor is idle, overload conditions, or future expansion, if applicable). The recommended cooling pump capacity margin is 10–25% (that is, pump rated capacity is 10–25% more than required normal flow).

Users should consider suitably sized pulsation vessels and correct any potential pulsation resonance in piping rather than using damping devices, such as orifices, choke tube and so on to dampen pulsation. Acoustic reviews should be performed during compressor system design to guarantee all anticipated combinations of pressures, speeds and load steps (including the use of flow-reduction steps that rely on unloaders, which can vary the compressor flow).

Pulsation limits are recommended around 85–95% of API 618 (Approach 3) limits to provide some margin (5–15% of API 618 limit values) to mitigate risk during construction and installation periods, and to cope with unanticipated deviations and problems. Similarly, pulsation vessels are generally fabricated before finalization of the piping-design-and-pulsation study, and enough margin should be provided to meet potential risks.

For nearly all applications, horizontal suction and horizontal discharge vessels are preferred. Long distances between vertical pulsation vessels and compressors increase the likelihood of pulsation problems.
Improving performance

The maximum predicted discharge temperature for any API 618 reciprocating compressor for CPI applications must not exceed 150ºC, and must not exceed 135ºC for hydrogen-rich service. In general, gas discharge temperatures below 118ºC tend to lead to longer life for the wearing parts.

When it comes to optimum pressure-drop values for pulsation dampeners and suppression devices, the pressure drop maximum is 1% of absolute pressure. For the intercooler, pressure drop around 0.70 bar or 2% of absolute pressure is recommended.

Readers should note that the use of orifice plates to dampen pulsation, especially on high-speed, single-act compressors (that is, those that compress gas on only one head of cylinder), can contribute to significant pressure drops.

To gain a better understanding of reciprocating compressor performance and track ongoing operation, the following performance curves should be developed:

• Suction pressure versus load
• Suction pressure versus flow
• Discharge pressure versus load
• Discharge pressure versus flow
• Suction pressure versus discharge pressure, per load step (that is, for each flow-reduction step using unloaders, typically 50%, 75%, 100% flow; 25% is rarely used because of the potential for reliability and load-reversal issues)
  

Such flow curves typically plot the minimum achievable flowrate to the maximum achievable flowrate in specified increment steps (for instance, in 10% steps). (Flow-versus-discharge-pressure plots of specific suction pressures may be an acceptable alternative when suction-pressure variations are limited). A review of the steepness of the proposed load curves can help the engineer to quickly identify which load curves (and where) are too steep. In these situations, small changes in pressure can have significant changes in load and flow. In general, compressors with steep load curves are hard to automate and tune. Thus, steep load curves usually indicate improper sizing of cylinders.

Optimum conditions

When scoping a reciprocating compressor system, it is absolutely necessary to have a minimum of two technically accepted proposals from qualified vendors. Small and medium compressors should be delivered fully fabricated as one skid-mounted package. Larger compressors are typically delivered as a prefabricated system (including the crankcase, distance pieces, and so on) with dismantled cylinders. Assembled cylinders are typically delivered to the site separately and installed later. It is common for the vendor to provide site-supervision work for cylinder installation at a negotiated lump-sum price.