IntroductionLarge-frame reciprocating com-

pressors have been used widely in oil and gas applications since the 1920s. As the large number of facili-ties utilizing this type of equipment has increased, so has the need for maximizing operational reliability while minimizing maintenance ex-penditures. Compressor analysis is a critical tool that provides a health condition assessment, ensures effi-cient and safe operation, accurately measures machinery performance, and provides a measurable econom-ic benefit for the operator.

This paper investigates excessive fuel consumption of compressor driv-ers caused by common compressor faults. Pressure versus volume (PV) analysis techniques will identify defi-ciencies, quantify fault severity, and will be used to estimate the resulting excessive fuel consumption. Empiri-cal fuel measurements of the driv-ers are analyzed before and after the fault correction and are used to calculate immediate economic sav-ings from repairs. Performance and capacity improvements are also ana-lyzed, providing a complete econom-

ic picture of maintenance and opera-tional payback.

BackgroundReciprocating compressor analy-

sis has been in use for many years, growing from the oscilloscope-based analyzers to modern digital portable and online systems. These tools have allowed analysts to evaluate com-pressor health, accurately measure performance and provide protection against potential catastrophic failures. However, the economic benefit of compressor analysis and condition-based maintenance is often difficult to quantify or is not utilized.

The measure of in-cylinder dy-namic pressure versus crank angle provides the cornerstone of com-pressor analysis. Often denoted as pressure versus time (PT), these curves can be converted into a pres-sure versus swept volume (PV) plot. By comparing actual versus theoreti-cal PV plots, leakage within the com-pressor cylinder can be identified, including suction/discharge valves, rings, and packing. These leakages are often confirmed and further dis-tinguished by vibration, ultrasonic and temperature readings, as well as numeric data derived from the pres-sure readings, such as flow balance and leak index. Pressure data are used to calculate compressor perfor-mance parameters, such as horse-power consumed, gas throughput, flow and efficiency. Finally, pressure

and vibration data play a critical role in identifying potential failure modes, including rod overload, lack of rod re-versal, and crosshead looseness.

A leak within a compressor cylin-der often results in the recirculation of gas. For example, if a discharge valve is leaking during the suction event, gas, which has already been compressed to discharge pres-sure, will leak back into the cylinder while it is at a lower pressure. Re-compressing the same gas results in less throughput and additional power requirements. The amount of throughput or flow can be accurately calculated by thermodynamic analy-sis of the cylinder PV parameters. The additional cost to recompress the gas can be estimated by exam-ining the ratio of gas moving into the cylinder versus the amount of gas leaving the cylinder (flow balance), and multiplying the percentage of this loss by the horsepower cost for the cylinder to compress gas. The resultant is termed “recirculation loss,” an estimate of the economic value of the cost of the leak, often expressed in $/day or $/year of com-pressor operation.

While the recirculation loss can pro-vide an estimate of the cost of cylinder leakage, to measure the actual cost, power consumption measurements are necessary. Since many high-speed compressors are driven by natural gas engines, measuring fuel flow and multiplying by fuel cost can

Economic Benefits Of Compressor Analysis > Improving gas flow through a compressor maximizes

value and revenueBy EDWARD B. FLANAGAN, PE

Edward B. Flanagan, PE, is the general manager at Windrock Inc. His area of expertise is systems analysis and de-sign, analog and digital design, instru-mentation and controls, and the appli-cation of instrumentation for machinery health and control.

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n Figure 1. Case 1: discharge valve leakage. The unit is a Caterpillar 3608 engine driving an Ariel JGD-4 three-stage compressor with a Windrock 6320/AP compressor analyzer. The application for Case 1 is natural gas gathering.

n Figure 2. Cylinder 4 PV with leaking discharge valve(s)(top).

n Figure 3. Cylinder 4 PT and ultrasonic traces confirming leaking discharge valves (bottom).

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n Table 1. Compressor report before repair. Note the low flow balance for cylinder 4 head end (red box) and the high rod loads for both stage-one cylinders (blue boxes).

Compressor Cylinder ID ihp @ rpm ihp (MMscfd) Capacity (MMscfd) Date Time

1 > Comp 1 H Pressure 198.0 @ 991.5 79.71 2.48388 9/8/14 16:01:48

2 > Comp 1 C Pressure 201.1 @ 991.2 79.01 2.54490 9/8/14 16:03:32

3 > Comp 2 H Pressure 314.3 @ 991.0 48.98 6.41673 9/8/14 16:16:55

4 > Comp 2 C Pressure 282.7 @ 992.0 47.40 5.96321 9/8/14 16:18:47

5 > Comp 3 H Pressure 198.7 @ 991.6 79.17 2.50968 9/8/14 16:06:57

6 > Comp 3 C Pressure 199.0 @ 986.9 78.88 2.52300 9/8/14 16:09:10

7 > Comp 4 H Pressure 365.0 @ 990.8 51.13 7.13873 9/8/14 16:12:39

8 > Comp 4 C Pressure 314.5 @ 987.6 45.16 6.96402 9/8/14 16:14:07

% Volumetric Efficiency

% Power/Valve Loss

% Flow Balance

Toe Pressure Com-

pressionRatio

Temperature (°F) Rod Load (%) Minimum Rod

ReversalDischarge Suction Discharge SuctionSuction/

DischargePd Ps Discharge Suction Tension

Com-pression

1 > 15.2 50.4 3.6 3.3 1.07 206.12 39.68 4.06 247.8 79.3 97.9 96.9 135 C

2 > 15.9 52.0 3.3 3.7 1.05 205.64 39.26 4.08 247.8 79.3 97.9 96.9 135 C

3 > 40.2 78.5 5.4 4.9 0.99 949.92 398.16 2.34 227.0 113.3 62.1 70.3 127 C

4 > 40.6 78.5 6.4 5.9 1.03 935.21 409.28 2.24 227.0 113.3 62.1 70.3 127 C

5 > 15.3 50.2 3.2 3.0 1.06 207.94 40.25 4.05 251.4 79.3 98.0 96.4 134 C

6 > 15.8 51.5 3.0 3.7 1.07 207.00 40.80 3.99 251.4 79.3 98.0 96.4 134 T

7 > 56.1 72.9 7.1 3.9 0.78 424.03 203.89 2.01 246.0 123.9 62.4 69.4 118 T

8 > 45.6 81.4 7.3 5.3 1.07 411.38 197.24 2.01 246.0 123.9 62.4 69.4 118 T

n Figure 4. Rod load for cylinder 3. Note the gas load is near the maximum tension and compression rating, and there is significant vibration knock at the crosshead.

n Figure 5. One of the two faulty discharge valves that were replaced.

calculate power consumption. The re-sult is expressed in $/day. When that number is divided by the actual com-pressed gas flow, the cost per flow or $/MMscfd is determined. By calculat-ing this value for a compressor with a known valve leak and then again after the valve is repaired, the actual fuel savings are the difference.

In addition to excessive fuel re-quirements, compressor faults often result in loss of throughput. If a com-pressor is incapable of providing its rated throughput capacity and that capacity is required (to be available for sale or because it limits a larger production process), this loss of flow can directly result in significant economic losses. For example, a natural gas producer may be able to sell gas for US$3.50/MMBtu. As-suming a heating value of 1024 Btu/scf (38,153 kJ/m3), that provides po-tential revenue of US$3580/MMscfd or US$1,308,000/MMscf/yr. If by re-pairing a known compressor fault, the producer is able to flow an ex-tra 1.0 MMscfd (0.028 x 106 sm3/d), the resulting sale of the gas is worth US$1.3 million per year.

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The following two examples illus-trate use of compressor analysis to identify faulty conditions and deter-mine the economic benefits of the re-pairs (Figure 1). Case 1 is based on analysis of the Figure 1 compressor and Case 2b is based on analysis of the Figure 9 compressor.

Analysis of the compressor showed significant discharge valve leak on cylinder 4 (stage 2), illustrat-ed by the PV curves in Figure 2. The head-end pressure trace (solid blue curve) does not follow the theoreti-

n Figure 7. Rod load for cylinder 3 after repair. Note that rod load is lower and the crosshead knock is gone as compared to Figure 4.

n Figure 6. PV after repair.

n Table 2. Compressor report after repairs. Note the improved flow balance (red box) and reduced rod loads (blue boxes) compared to Table 1.

Compressor Cylinder ID ihp @ rpm ihp (MMscfd) Capacity (MMscfd) Date Time

1 > Comp 1 H Pressure 177.9 @ 988.2 74.63 2.38436 9/8/14 18:53:58

2 > Comp 1 C Pressure 186.0 @ 993.2 72.88 2.55269 9/8/14 18:52:55

3 > Comp 2 H Pressure 318.2 @ 991.3 46.36 6.86408 9/8/14 18:44:02

4 > Comp 2 C Pressure 286.4 @ 990.6 44.57 6.42670 9/8/14 18:45:13

5 > Comp 3 H Pressure 184.9 @ 996.0 75.84 2.43855 9/8/14 18:55:13

6 > Comp 3 C Pressure 185.8 @ 990.8 71.93 2.58329 9/8/14 18:56:11

7 > Comp 4 H Pressure 314.0 @ 984.5 53.18 5.90486 9/8/14 18:48:39

8 > Comp 4 C Pressure 308.6 @ 988.6 53.57 5.76106 9/8/14 18:50:16

% Volumetric Efficiency

% Power/Valve Loss

% Flow Balance

Toe Pressure Com-

pressionRatio

Temperature (°F) Rod Load (%) Minimum Rod

ReversalDischarge Suction Discharge SuctionSuction/

DischargePd Ps Discharge Suction Tension

Com-pression

1 > 16.7 48.4 2.9 2.6 0.98 181.65 36.96 3.80 236.5 83.4 89.0 86.8 154 C

2 > 17.6 52.4 2.6 3.4 1.00 181.29 36.94 3.80 236.5 83.4 89.0 86.8 154 C

3 > 41.6 79.9 5.3 5.4 1.01 959.75 418.54 2.25 218.6 111.2 69.0 76.4 162 C

4 > 42.9 79.9 6.8 6.5 1.03 941.17 431.14 2.14 218.6 111.2 69.0 76.4 162 C

5 > 16.8 48.4 2.5 2.8 0.97 183.49 37.38 3.81 233.2 80.0 88.6 87.4 154 C

6 > 17.8 52.1 2.7 3.2 1.00 181.22 37.60 3.75 233.2 80.0 88.6 87.4 154 C

7 > 34.8 71.1 6.0 4.1 1.02 440.03 174.37 2.41 238.9 118.9 82.3 85.0 154 C

8 > 34.9 73.6 4.9 3.9 1.04 439.58 171.39 2.44 238.9 118.9 82.3 85.0 154 C

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cal curve (dashed blue line). With the measured pressure coming up to dis-charge pressure too quickly and back down to suction pressure too slowly, a classic discharge leak pattern is identified. Figure 3 confirms the leak with ultrasonic measurements. As indicated, the ultrasonic measure-ments should be quiet (thin line) dur-ing the head-end suction event. The excess noise in this area is indicative of a discharge valve leak. Also note in Figure 2 that the automated Leak Index tool also identifies the dis-charge leak.

The calculated performance pa-rameters also confirm a discharge valve leak with a low flow balance (below 0.95) as shown in Table 1. The low flow balance indicates that more gas is flowing into the cylinder than is leaving the cylinder over the course of each stroke.

The leaking discharge valve in the cylinder 4 head end is allowing dis-charge gas to recirculate back into the cylinder during the suction stroke. The poor efficiency in stage 2 caused stage 1 cylinders to perform a higher share of the work for compressing the gas. Because of this imbalance, cyl-inders 1 and 3 are dangerously close to rod overload and have caused a significant vibration knock in the crosshead when the piston changes from tension to compression. The overload can be seen numerically in Table 1 and graphically in Figure 4, which also shows the crosshead vi-bration knock.

With the compressor analysis pin-pointing discharge valve leakage on the head end of cylinder 4 and a potential for catastrophic failure of cylinder 3, the maintenance crew opened cylinder 4 to confirm and replace the faulty discharge valves (Figure 5).

After the repairs were complet-ed, the actual PV and theoretical PV curves lined up very closely, as shown in Figure 6. The flow balance improved from 0.78 to 1.02 for the cylinder 4 head end, as shown in Table 2. The same table also shows a reduced rod load for the first-stage cylinders. Figure 7 shows the re-

n Table 3. Compressor performance and fuel flow economics.

Parameter Source Before Repair After Repair Difference

Compressor Power (hp)

Windrock MD Software

2182 2065 -5.40%

Gas Flow (MMscfd)

Windrock MD Software

12.38 13.29 7.35%

Compression Ratio

Windrock MD Software

17.9 18.7 4.50%

Fuel Flow (scfh)

Measured 17,285 16,983 -1.70%

hp/MMscfd Calculated 176.3 155.4 -11.90%

Fuel Flow/hp Calculated 7.92 8.22 3.80%

US$/hp-hr (@ $4/Mscf)

Calculated $0.031 $0.032

US$/d Calculated $1,660 $1,630 -1.80%

US$/MMscfd Calculated $134.10 $122.65 8.5% ($11.45)

n Figure 8. Theoretical losses.

n Figure 9. Case 2: suction valve blockage. The unit is a Caterpillar 3516 engine driving an Ariel JGT-4 three-stage compressor with a Windrock 6320/PA compressor analyzer. The application for Case 2 is natural gas gathering.

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duced rod load graphically, as well as the elimination of the crosshead knock that was seen in Figure 4.

Economic analysis for discharge valve leak

The leaking discharge valves in the head end of cylinder 4 (stage 2) caused discharge gas to re-enter the cylinder during the suction stroke. This gas expands during suction and is recompressed during the compres-sion stroke. Recirculating this gas causes additional work by the driver. Windrock MD software uses the flow balance to estimate the percentage of recirculated gas and multiplies that by the horsepower required for the head end of cylinder 2. Using a stan-dard horsepower cost of $0.032/hp-hr (based on a fuel cost of $4/Mscf), the software calculated a recircula-tion loss of $112/d or nearly $41,000/yr. Figure 8 represents the excess power required to recompress gas that leaked back in through the faulty discharge valves.

n Table 4. Compressor report before repair. Note the high power losses for the suction events for cylinder 2 (red boxes).

Compressor Cylinder ID ihp @ rpm ihp (MMscfd) Capacity (MMscfd) Date Time

1 > Comp 1 H Pressure 47.8 @ 1382.8 64.77 0.73760 9/14/14 12:46:15

2 > Comp 1 C Pressure 107.0 @ 1381.2 72.46 1.47653 9/14/14 12:47:22

3 > Comp 2 H Pressure 187.2 @ 1382.8 70.18 2.66791 9/14/14 12:52:53

4 > Comp 2 C Pressure 157.1 @ 1382.6 63.34 2.48013 9/14/14 12:54:14

5 > Comp 3 H Pressure 49.5 @ 1380.8 72.08 0.68704 9/14/14 12:50:25

6 > Comp 3 C Pressure 112.2 @ 1381.7 69.93 1.60473 9/14/14 12:51:19

7 > Comp 4 H Pressure 179.1 @ 1378.6 77.46 2.31222 9/14/14 12:55:19

8 > Comp 4 C Pressure 159.4 @ 1380.2 74.86 2.12891 9/14/14 12:55:54

% Volumetric Efficiency

% Power/Valve Loss

% Flow Balance

Toe Pressure Com-

pressionRatio

Temperature (°F) Rod Load (%) Minimum Rod

ReversalDischarge Suction Discharge SuctionSuction/

DischargePd Ps Discharge Suction Tension

Com-pression

1 > 11.3 27.0 8.7 4.0 1.04 117.40 29.70 2.98 220.0 68.0 50.7 49.2 126 T

2 > 23.1 57.3 10.0 6.8 1.04 116.50 28.35 3.05 220.0 68.0 50.7 49.2 126 T

3 > 34.1 75.8 9.1 15.7 1.02 877.07 304.80 2.79 241.8 85.0 53.5 57.6 145 T

4 > 35.7 74.2 9.3 13.8 1.04 863.56 326.40 2.57 241.8 85.0 53.5 57.6 145 T

5 > 10.0 24.9 5.7 3.2 1.02 123.70 29.70 3.12 218.2 71.2 50.5 49.7 125 T

6 > 24.9 58.7 9.4 6.2 1.00 117.60 29.41 3.00 218.2 71.2 50.5 49.7 125 T

7 > 29.7 75.7 9.9 7.8 1.04 386.47 108.43 3.26 256.1 79.1 61.0 62.8 137 C

8 > 28.8 74.5 8.7 7.4 1.08 381.39 110.83 3.16 256.1 79.1 61.0 62.8 137 C

Stage Capacity (MMscfd)

1 4.5059

2 4.4411

3 5.148

n Figure 10. PV curves before repair.

Toe Pressure Compression

Pd Ps Ratio

117.40 29.70 2.98

116.50 28.35 3.05

877.07 304.80 2.79

863.56 326.40 2.57

123.70 29.70 3.12

117.60 29.41 3.00

386.47 108.43 3.26

381.39 110.83 3.16

n Table 5. Cylinder pressures.

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By evaluating the fuel flow read-ings before and after the repair, a more precise economic analysis can be developed, as shown in Table 3.

After the repair, the compressor moves more gas with less horsepower. Assuming this compressor is running 24/7 at a flow rate of 13.29 MMscfd (0.376 x 106 sm3/d), a fuel savings of US$55,500 per year is realized with the replacement of the two head-end discharge valves on cylinder 4 (stage 2). In addition to these fuel savings, there are intangible savings from low-ering the rod loading and distribut-ing the load among the stages more evenly, resulting in less component

wear and avoidance of potential cata-strophic failure.

Initial analysis revealed a relatively healthy compressor with little recircu-lation. In contrast to the leaking dis-charge in Case 1, the actual PV plots follow closely with the theoretical curves, the flow balances are close to 1.0 (indicating a healthy cylinder), and the ultrasonic data does not show any cylinder leakages. Upon closer inspection of the compressor report data in Table 4, a higher than normal power loss is found during the suction event for cylinder 2 (stage 3). Power loss represents the energy required to pull the gas into the cylinder from

the suction bottle. It can have several components: the power required to open and flow gas through the suction valves and pressure losses through the nozzles and any restrictions, such as orifice plates.

As seen in Figure 10, the PV curve shows very high suction horsepower losses, confirming the data in the compressor report. The values of 15.7 and 13.8% suction valve horse-power losses computes to 50 hp (37.3 kW) additional load required to pull in gas to this compressor cylin-der. A more typical figure would be 7 or 8%. The cause of this excess power requirement could be clogged suction valves. Another possibility would be a clogged screen in the nozzle of this machine, or even a clogged cooler between the second and third stages. The odd pointed down shape of the PV toe indicates excessive drawing down of the bottle pressure, representative of a block-age between stages.

From close examination of the com-pressor report, there is a pressure drop from the first to the second stage of 7 psi (0.48 bar), and a drop from the second to the third stage of 70 psi (4.83 bar). This is also indicative of a blockage. The compression ratios across the stages are not balanced, resulting in an uneven distribution of the work across the compressor.

The third-stage suction valves were pulled and found to be signifi-cantly clogged with a white crystal-line substance, as can be seen in Figure 11.

After the repair, additional analysis data were gathered on the machine to quantify the improvement made by changing the clogged suction valves.

n Figure 11. Clogged suction valves.

n Figure 12. PT curves before and after repairs.

Compressor Cylinder ID ihp @ rpm ihp (MMscfd) Capacity (MMscfd) Date Time

BEFORE

3 > Comp 2 H Pressure 187.8 @ 1382.8 70.93 2.64709 9/14/14 12:52:53

4 > Comp 2 C Pressure 157.5 @ 1382.6 63.72 2.47112 9/14/14 12:54:14

AFTER

3 > Comp 2 H Pressure 199.9 @ 1382.5 52.55 3.80416 10/28/14 12:28:43

4 > Comp 2 C Pressure 165.5 @ 1379.8 49.88 3.31746 10/28/14 12:29:28

n Table 6. Measured performance data before and after valve repair.

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x 106 sm3/d) of gas for this unit was costing the user US$2.6 million per year in potential revenue!

ConclusionsCompressor analysis is an invalu-

able tool that provides direct insight into the health and performance of reciprocating compressors. As part of a condition-based maintenance program, reciprocating analysis has proven to reduce maintenance costs and protect against catastrophic ma-chinery failures.

The economic benefit of compressor analysis can be quantified by theoreti-cal models or more directly by examin-ing the driver cost before and after re-pairs are performed. The fuel savings from a typical discharge valve leak are significant and can quickly provide a return on investment for an analy-sis program. By improving gas flow through a compressor, an operator can maximize the economic value of a reciprocating compressor and greatly improve revenue generation. CT2

lowest cost. Breaking this down to cost, and using a driver cost of $0.032/hp-hr, the following conclusion can be made.

Before repair: $18,200/MMscf for one year, or to

pump 7 MMscfd (0.198 x 106 sm3/d) for one year, $127,000.After repair:

$13,900/MMscf for one year, or to pump 7 MMscfd for one year, $97,000.

To pump the same amount of gas through this cylinder, the clogged valves were costing the owner US$30,000 per year.

More importantly, the clogged valves on this machine were throttling capacity to 5 MMscfd (0.142 x 106 sm3/d) whereas the machine should have been capable of 7 MMscfd (0.198 x 106 sm3/d). This lost capacity resulted in lost revenue for the produc-er. As previously mentioned, assum-ing a sale price of US$3.50/MMBtu for the gas, every MMscf of gas is worth US$1.3 million. Therefore, the reduced capacity of 2 MMscfd (0.057

Figure 12 illustrates the pressure waveforms before and after the suc-tion valves were replaced. It is visual-ly apparent from the waveforms that the clog caused a severe restriction and pressure drop inside the cylinder during suction, thus lowering the final discharge capacity of the cylinder.

Economic analysis of restricted suction valve

From theoretical models, repair of this issue resulted in about a 25 hp (18.64 kW) reduction in load for the same amount of production. Using a driver cost of US$0.32/hp-hr, this reduction in horsepower would result in a savings of approximately US$7000 per year.

However, by examining the data in Table 6, the stage capacity increased from 5.12 to 7.12 MMscfd (0.145 x 106 to 0.202 x 106 sm3/d) after the repair. This is a 39% improvement in flow. The horsepower before was 345 and 365 after — only a 5% increase.

The goal of analysis is to help op-erators compress and pump gas at the

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