The Various Loads Used to Rate Reciprocating Compressors (Part Two)


Written by:
K.E. Atkins, Martin Hinchliff and Bruce McCain
 A Note from Robert X. Perez:

Welcome back to Compressor University!

Here is the second installment from the Atkins, Hinchliff and McCain article. This month they continue their review of reciprocating compressor load limit definitions.

Robert X. Perez

History of "Rod Loads"

The 1st Edition of API-618 was published in 1964 (34 pages). It included no definition of what the term "Rod Load" meant. However, the data sheets did call for the compressor manufacturer to specify the "Max Allowable Rod Loading" and "Rated Rod Loading." So the definition of what that meant was left up to the compressor OEM.

In the 1963 edition of the Ingersoll-Rand (IR) frame ratings guide, the piston loads were defined. This document stated that piston load is frequently referred to as "rod load," which is a misnomer as it implies that the piston rod is the only limit in the establishment of a compressor load rating. It defined the piston load as the nominal pressure at the cylinder flange times the area of the piston.

These loads were easily calculated from simple equations (available in Part One). It went on to say that the actual rod load would include the effect of inertia and valve losses, but these effects were considered in the piston rod load ratings, i.e. the piston rod load ratings were necessarily conservative. This approach served the industry well, but perhaps resulted in "over-designed" machinery.

This was before the advent of electronic calculators and digital computers, so combined rod load was tedious to calculate. The practice at the time was to look at and report simply the nominal gas load only with no valve losses or inertia loads considered. On rare occasions if it was judged necessary due to a combination of high gas loads, high inertia forces and high volumetric efficiency (which can cause the gas load and inertia load to be additive), a manual calculation of combined rod load (gas + inertia + valve losses) would be done.

This would consist of drawing a PV card including valve losses, using a planimeter and slide rule to determine area (horsepower) and gas pressures at discrete degrees of rotation increments. Then inertia forces were calculated at each point and added to determine the combined rod load at the crosshead pin, forces in the connecting rod and crankshaft, and torque on the crankshaft. For a six-throw compressor it would typically require six engineers (one cylinder each) and one week to perform this task.

The 2nd Edition of API-618 was published in 1974 (39 pages). The committee pushed the compressor manufacturers to advise how rod loads were calculated and ensure that everyone would calculate rod loads the same way. This established the term "allowable rod load" and "actual rod loading." The actual rod load was defined as the force due to the differential pressure across the piston plus the inertia of the reciprocating parts transmitted through the piston rod. It also stated that the actual rod load calculated on the basis of cylinder relieving pressure (RV setting) shall not exceed the vendor's maximum allowable rod load.

By this time mainframe computers and programmable calculators were widely used. This allowed for more precise engineering calculations and the elimination of some of the conservatism in the design process. Practice was to calculate compressor performance and "gas load" using a programmable calculator, since the computations were relatively simple. Basic compressor sizing and feasibility studies used these methods.

The final performance, including actual rod load (combined rod load), was obtained using mainframe computers (punch cards, overnight batch processing, etc.). Gas loads were still calculated and reported based on nominal cylinder flange gas pressures, but actual rod load included the effect of valve pressure drop and inertia loads. There was variability between various users and OEMs on the reference points used for the calculation of combined rod load. At IR and Worthington, the reference point was the crosshead pin, so all inertia outboard of the pin bearing was included in the combined rod load calculation. There was also inconsistency over the relief valve pressure. Some users and OEMs (including IR) used the final relief valve pressure rather than each stage RV setting.

In the 3rd Edition (1986), API-618 grew to 111 pages. The term Maximum Allowable Combined Rod Load (MACRL) was defined. The combined rod load was defined as the algebraic sum of the differential gas pressure on the differential piston area plus the inertia force. The reference point for inertia loads was defined as being at the crosshead pin. Additionally API established a minimum rod load reversal criteria (to ensure proper lubrication, 3 percent and 15 degrees), but that issue is outside the scope of these articles. Gas load still was reported based on nominal cylinder flange pressures and the relief valve pressure calculation was still inconsistent.

In the 4th Edition (1995) API-618 was at 166 pages. The calculation of rod load was defined much more precisely. The terms Max Allowable Continuous Combined Rod Load (MACCRL) and Max Allowable Continuous Gas Load (MACGL) were established. This was the first time that load limits based on running gear and load limits based on the stationary components were explicitly separated in the specification. Combined rod load was defined the same way as the 3rd Edition but with the clarification that it was to be at the crosshead pin and only the component in the direction of piston motion was included. Note that the load in the connecting rod is higher due to geometry.

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