Integration of Special Purpose Reciprocating Compressors into a Process

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Engineering Design Guide: GBHE-EDG-MAC-1033

Integration of Special Purpose Reciprocating Compressors into a Process Information contained in this publication or as otherwise supplied to Users is believed to be accurate and correct at time of going to press, and is given in good faith, but it is for the User to satisfy itself of the suitability of the information for its own particular purpose. GBHE gives no warranty as to the fitness of this information for any particular purpose and any implied warranty or condition (statutory or otherwise) is excluded except to the extent that exclusion is prevented by law. GBHE accepts no liability resulting from reliance on this information. Freedom under Patent, Copyright and Designs cannot be assumed.



Engineering Design Guide: Integration of Special Purpose Reciprocating Compressors into a Process Service

CONTENTS SECTION 1 SCOPE 1 2 CHOICE OF COMPRESSOR TYPE 2

2.1 Parameters

2.2 Preliminary Choice of Machine Type

2.3 Review of Other Types of Compressor 3 CHOICE OF NUMBER OF COMPRESSORS 3

3.1 Influence of Reliability Classification

3.2 Driver Considerations

3.3 Deterioration of Standby Machines 4 EFFECTS OF PROCESS GAS COMPOSITION 4

4.1 Particulate Contamination

4.2 Droplets in Suspension

4.3 Polymer Deposit 4.4 Molecular Weight Variation



4.5 Compressibility Variation

4.6 Gas Dryness

4.7 Gas Solution in Lubricating Oil for Cylinder and Gland 5 THROUGHPUT REGULATION 5

5.1 Inlet Line Throttle Valve

5.2 Inlet Line Cut-off Valve

5.3 Compressor Inlet Valve Lifter 5.4 Clearance Volume Variation

5.5 Speed Variation

5.6 Bypass

5.7 Hybrid Regulation Systems

6 PRINCIPAL FEATURES 6

6.1 Calculate Discharge Gas Temperature

6.2 Choice of Number of Stages

6.3 Configuration 6.4 Valve Operation Limit on Piston Speed

6.5 Limits for Mean Piston Speed

6.6 Estimation of Volumetric Efficiency

6.7 Estimation of Crankshaft Rotational Speed

6.8 Calculation of Piston Diameter

6.9 Choice of Number of Cylinders



7 DRIVER TYPE 7

7.1 Electric Motors

7.2 Steam Turbines 7.3 Special Drivers

8 VESSELS 8 APPENDICES A RELIABILITY CLASSIFICATION B CONDITIONS FOR LUBRICATED CYLINDERS AND GLANDS C ESTIMATE OF LUBE OIL CONTAMINATION OF PROCESS GAS D INFLUENCE OF GAS COMPOSITION AND MACHINE CONSTRUCTION

ON FILLED PTFE PISTON RING SEALS E LIMITS ON GAS TEMPERATURES FIGURES 1 SELECTION CHART 2 DESIGN SEQUENCE 1 - ESTIMATE NUMBER OF STAGES 3 DESIGN SEQUENCE 2 - ESTIMATE CYLINDER SIZES DOCUMENTS REFERRED TO IN THIS ENGINEERING DESIGN GUIDE



1 SCOPE This Engineering Design Guide covers the estimation of the principal features of a reciprocating gas compressor to obtain a preliminary specification for enquiries to be sent to appropriate vendors. It applies to compressors as defined in the GBHE-EDP-MAC-3301 Series, and is also an essential preliminary step for a compressor in Group 1 whose final duty and specification is negotiated with the chosen supplier. It is not intended to cover service air compressors where the discharge pressure is less than 9 bar gauge, nor compressors of the diaphragm type. 2 CHOICE OF COMPRESSOR TYPE 2.1 Parameters (a) Calculate maximum actual volume flow at inlet as: Q i = W x Z i x R x Ti m3/s 100 x M x Pi

W - Mass flowrate kg/s

Ti - Inlet gas absolute temperature K Pi - Inlet gas absolute pressure bar abs

Z i - Gas compressibility corresponding to Pi , Ti

R - Gas constant of 8.3143 kJ /kg-mol K

M - Gas molecular weight

Note that W, T i , M and Pi form that set of self consistent values from the Process Data Sheet which gives the maximum value of Q i.



(b) Calculate the isothermal head H t as: Ht = Zi x R x Ti Loge Pd kJ/kg M Pi Where Pd = maximum discharge pressure bar abs (c) Estimate the power required as:

E = 1.6 W x Ht kW 2.2 Preliminary Choice of Machine Type Enter Fig. 1 Selection Chart (To confirm the preliminary choice of a

reciprocating machine.)



2.3 Review of Other Types of Compressor Other types of compressor are worth investigation when the following criteria are satisfied. (a) Centrifugal compressors should be considered when the minimum

notional volume flow, measured at conditions of discharge pressure but inlet temperature, exceeds 0.25 m3 /s.

(b) Oil-free screw-compressors should be considered when:

(1) The volume flow at inlet conditions of temperature and pressure is within the range 0.6 - 5.5 m3 /s

(2) The maximum discharge pressure is 40 bar g

(3) The maximum differential pressure across any stage up to a

discharge pressure of 16 bar is 8 bar

(4) The maximum differential pressure across any stage above an inlet pressure of 10 bar is 16 bar

(5) The maximum compression ratio of any stage is 4.5

(c) Oil-free Roots compressors should be considered when:

(I) The maximum differential pressure is 1.4 bar

(2) The inlet actual volume flow is within the range 0.1 to 1.0 m3 /s

(d) Consider hybrid configurations having a rotary or centrifugal compressor

feeding one or more reciprocating compressors,

Where:

(I) The actual volume flow at inlet conditions exceeds 4.5 m3/s

(2) The notional volume flow (measured at discharge pressure but at inlet temperature) is less than 0.25 m3/s.



3 CHOICE OF NUMBER OF COMPRESSORS 3.1 Influence of Reliability Classification The classifications are defined in Appendix A. Class I This degree of reliability cannot be reached by commercially available reciprocating compressors: consequently two 50% duty or three 33% duty compressors running in parallel should be specified, depending on the acceptable degree of process upset and the desired range of capacity regulation. Class 2 The first choice is a single 100% duty compressor. The alternative configuration of two 50% duty compressors may be chosen on comparison of installed capital cost or on the desired range of capacity regulation. Class 3 This classification does NOT apply to reciprocating compressors because it is not feasible to permit significant deterioration by corrosion or by particle-induced wear. Classes 4, 5 and 6 For most duties the use of more than 3 machines is unnecessary; more than 4 machines is uneconomic unless an existing plant is being extended. 3.2 Driver Considerations Low power steam turbines are relatively expensive and inefficient. More than two compressors in parallel should not be used when one or both are driven by steam turbines of less than 1.5 MW rating.



3.3 Deterioration of Standby Machines In choosing the number of installed machines remember that standby machines will require special measures to withstand deterioration during standstill periods. 4 EFFECT OF PROCESS GAS COMPOSITION 4.1 Particulate Contamination The gas should not contain hard crystalline particles larger than 3µm. Most process gases are sufficiently clean provided that the compressor inlet is preceded by a condensing cooler and an effective droplet separator. Ostensibly clean dry gases need careful evaluation: permanent inlet gas filters are normally needed where products of corrosion of preceding equipment can be carried forward or where catalysts dust may be present. 4.2 Droplets in Suspension Liquid droplets that do not evaporate during the compression phase significantly affect valve life. Specify a droplet separator whose effective cut-off size rating is less than 10µm. When evaporation occurs, e.g. water droplets in air, the separator effective cut-off size rating may be increased to 20 µm. 4.3 Polymer Deposit Polymer can normally be removed by solvent injection whilst the compressor is being barred over. This requires provision of: (a) An injection system, atomizing the solvent (b) Motorized continuous barring. (c) An effective drain system, with vertically downward discharge cylinder

connections.



4.4 Molecular Weight Variation Such variation significantly affects: (a) The volumetric efficiency of labyrinth compressors, because piston bypass

increases as molecular weight decreases. (b) The gas forces applied to the valves. Verify the range of gas molecular weights over which the compressor will operate, especially for mixtures containing hydrogen or helium. Ranges exceeding 2:1 require close limits on design gas velocity through valves; those exceeding 3:1 may reduce valve life expectancy. 4.5 Compressibility Variation Operation near the saturation line and particularly near the critical point requires intensive assessment of start-up conditions and the likely fluctuation in process pressure and temperature. Two-phase operation is unacceptable. Operation with the gas in the quasi-liquid condition requires a comparison of mean piston speed and mean gas speed through the valves with those values appropriate to liquid pumps. 4.6 Gas Dryness Gases derived from cryogenic separation, or allied plants which contain no water vapor or oxygen, affect the life expectancy of piston rings for both oil-lubricated and oil-free machines. (a) For oil-lubricated machines scuffing may take place in reducing

atmospheres. See clause 6.5.1. (b) For oil-free duties, labyrinth compressors should be selected unless the

vapor content exceeds the limiting values given in Appendix D, when PTFE-based piston rings may be used.



4.7 Gas Solution in Lubricating Oil for Cylinder and Gland Estimate the viscosity change due to dilution of the oil by the gas, using the method given in Appendix B, for each stage. Give the resulting choice of oil grade in the enquiry specification. 5 THROUGHPUT REGULATION Reciprocating compressors run at constant speed give roughly constant volume flow. 5.1 Inlet Line Throttle Valve Specify this method for: (a) Exhauster duties, where the inlet pressure is sub-atmospheric and the

flow range is normally less than 2:1 Greater flow ranges are practicable but may require more than one stage to comply with the limit for gas discharge temperature.

(b) Any compressor intended to operate at approximately constant mass flow

where there is a major variation in the inlet gas pressure or temperature. 5.2 Inlet Line Cut-Off Valve This method gives regulation over the whole 100% flowrange. The compressor operates cyclically between maximum flow and zero flow: the flow variation is smoothed by providing a large air receiver. The delivery pressure fluctuates over a small range (not less than 0.3 bar). More than one compressor may be connected to the receiver, whereupon: (a) The pressure differential becomes [0.7 + 0.02 P] bar (b) Each compressor should be subject to a predetermined timed shut-down

sequence, usually after idling for 20 minutes. If there is no operator attendance the start/stop initiation should be automatic.

The cycling operation has inherent power losses at the start and finish of each cycle: this gives a lower limit to the period of operation and consequently to the size of the receiver. As a first estimate take the minimum reservoir volume as:



60 x Q m3 P x dP

where Q is the compressor volume flow

(at normal conditions of 1013 millibar and 0°C) Nm3/s

P is the mean discharge pressure bar abs dP is the chosen differential pressure band for the receiver bar

5.3 Compressor Inlet Valve Lifter Proprietary valve arrangements are available which lift the inlet valve plates during the nominal compression stroke of the piston so that the cylinder gas content is returned to the inlet line. This arrangement may be used for 'ON-OFF' action, giving the cyclical variation described in Clause 5.2 and subject to the same requirements for receiver size. Process duties may be satisfied by stepwise regulation where independent valve lifters are provided on the outer ends of double-acting cylinders. A compressor having two cylinders operating in parallel can have 3-step regulation of 50, 75 and 100% flowrate (regulation at 0 and 25% is not then practicable). 5.4 Clearance Volume Variation This method is efficient but cannot be used for gases that may lay down deposits. Note that it increases the cylinder gas discharge temperature. Continuous regulation over a range of about 20% is obtained by an auxiliary piston varying the volume of a chamber connected to the cylinder so that effective clearance volume of that end of the cylinder is continuously varied. Stepwise regulation is obtained by 'ON-OFF' valves isolating bottles connected to the cylinder.



5.5 Speed Variation With steam turbine or other special drivers it is possible to vary the speed and so vary the flowrate. Mechanical design difficulties typically limit such variation to less than 15%. 5.6 Bypass This method has the advantages of: (a) Quick response (b) Continuous variation (c) Simplest compressor construction The disadvantage is that the power efficiency falls as the delivery flow is reduced. The two principal arrangements are: (a) Bypass across whole compressor

This arrangement is normally employed for starting and shutdown purposes: it can cover a range of 0 to 100% capacity.

The bypass flow connection should be taken from the catchpot following the after-cooler; otherwise a bypass cooler may be required. Note that the Joule-Thompson effect normally cools the gas.

(b) Bypass across Stage 1 only

Full bypass effectively removes stage 1; consequently for a constant final discharge pressure the compressor is provided with one extra stage. The effective flowrange is then approximately the same as the rated compression ratio across stage 1.

The bypass should be fitted between the inlet catchpots to stage 1 and stage 2. The stage 1 intercooler then keeps the inlet gas temperature to stage 1 nearly constant.



5.7 Hybrid Regulation Systems All systems which reduce the capacity of only the first stage increase the compression ratio across the last stage (when the final discharge pressure is constant). Consider provision for either an additional stage or for a regulation system on the intermediate and final stages. 6 PRINCIPAL FEATURES 6.1 Calculate Discharge Gas Temperature. Compression is assumed to follow the pseudo-polytrophic relations: P x V n = constant

m-I = constant

T x P m

Where, in general, n > m Provided that the working range of gas conditions is far removed from the critical point for the gas, it is sufficiently accurate to take:

M = (1 + α)ᴕ 1 + α x ᴕ where ᴕ is the ratio of specific heats (Cp/Cv)

α is the performance factor For typical process compressors, α is small because it is the residual effect of two opposing influences; firstly the heat loss due to cylinder cooling, secondly the heat gain from mixing the charge gas with gas retained in the clearance volume of the cylinder from the previous compression stroke. As a first estimate take α as zero.



After the preliminary assessment of the compressor features, make a second estimate from:

α = 150 - 0.4 (1 - ᵑv) D x N1/2

Where D is the cylinder bore in mm. FIG 2 DESIGN SEQUENCE I ESTIMATE NUMBER OF STAGES



6.2 Choice of Number of Stages Determine the minimum number of stages from the following consideration: (a) The limiting discharge temperature from any stage, with the inlet

temperature at the highest value expected for summer ambient conditions and the pressure ratio at the corresponding maximum operating value. See Appendix E.

The stage inlet gas temperatures may be determined by the specified margin above dewpoint, when process or metallurgical reasons dictate dry operation.

(b) Minimum power demand at normal operating conditions. This is obtained

by equalizing the enthalpy change per stage, which is approximately obtained by making the compression ratio the same for each stage.

(c) Throughput regulation needs. (d) The minimum and maximum pressures of any side streams.



FIG 3 DESIGN SEQUENCE 2 ESTIMATE CYLINDER SIZES



6.3 Configuration Horizontal-opposed cylinder configurations are required for lubricated machines rated above 600 kW. Vertical cylinders are preferred for oil-free service generally and should be used for labyrinth compressors. Compressors of capacity less than 0.5 m3/s may have two cylinders arranged as a 'V'. Exhausters have a large first stage cylinder which is best arranged vertically, with a horizontal second stage forming an 'L'. 6.4 Valve Operation Limit on Piston Speed The limiting mean gas velocity through the valves of one stage is u m/s, where: U = 45 Pi

(r – 1) 3/8 ρ ½ r

ρ = Gas density at inlet stage conditions kg/m3

r = stage compression ratio (r > 1.10) Pi = absolute pressure at stage inlet bar abs

For a given cylinder design, u is related to the mean piston speed (U). As a preliminary estimate take: U < 0.42 U½ x ΔP1/4 m/s Where P is the differential pressure across the cylinder bar Note that on a multiservice compressor having one cylinder devoted to a booster or circulation duty where r <1.30, the crankshaft may be specified to have one short throw to give a lower piston speed for that cylinder.



6.5 Limits for Mean Piston Speed The criterion for these limits is the life expectancy of rubbing elements, viz. piston rings and packings, cylinder liners, motion work. 6.5.1 Oil Lubricated Cylinders For clean non-corrosive gases containing sufficient oxygen and water vapor (> 100 ppm) to ensure the renewal of oxide films on rubbing parts, the common materials of cast iron and bronze are satisfactory for speeds up to the limit imposed by inertia loading and alignment, which is currently:

Reliability Class

Mean Piston Speed m/s

2 & 4 4.5 5 & 6 5.2

Process gases affecting the lubricant or requiring other materials may impose a lower limit. Current practice is to add oxygen (air) and water vapor to the process gas or to the cylinder oil in order to prevent scuffing. If this cannot be done then the compressor should be compared with a reference installation having an IDENTICAL duty. 6.5.2 Oil-free Cylinders Using PTFE Based Rings Ring life expectancy is sensitive to alignment and loading as well as to the nature of the gas. This leads to the following criteria:

(a) In process gas compressors operating at constant conditions, the limiting speed decreases as the pressure differential across the piston increases. The effect is mitigated by increasing the number of rings.



Where ΔP is the differential pressure bar

across the piston Segmental packing for piston rod seals should comply with the same rules providing that the variation in pressure differential across the packing is sufficiently great. For a preliminary assessment take: ΔP s > 0.3 Ps - Po Where ΔP s is the variation in pressure bar Within the seal box Ps is the nominal pressure bar abs

within the sealbox during the piston rod instroke

Po is the nominally constant bar abs pressure at the seal exhaust vent.



(b) For compressors operating with throughput regulation by cyclic idling, the rings are subjected to alternate periods of loaded and unloaded operation. Current practice is to limit the mean piston speed to 82% of the values given in clause 6.5.2(a).

(c) Oil-lubricated PTFE-based piston rings may be used for piston speeds up

to 4.0 m/s. 6.6 Estimation of Volumetric Efficiency For each stage:

ᶯv = 1 - r-1 0.84 – 0.16 x r 1/n* 32 x δ Where

ᶯv is the fractional volumetric efficiency. r is the pressure ratio across the stage n* is the pseudo-polytrophic index for expansion

For the first estimate of ᶯv take the value of n* as 0.94δ

where δ is the ratio of specific heats Cp/Cv

6.7 Estimation of Crankshaft Rotational Speed

Upper bound of rotation speed: 2.8 U3 ½ r/s

ΔP x Q i



Where Q i is the actual inlet volume flow for one cylinder m3/S ΔP is the differential pressure across that cylinder bar U is the mean piston speed m/s Over-riding limits on this upper bound are: - For PTFE based piston rings 47 r/s U - For carbon piston rings 37 r/s U 6.8 Calculation of Piston Diameter For a single cylinder, double-acting, D = 20 x V I ½ x 104 mm ᶯv x π x S x N Where Vi is the actual volume flowrate for the conditions of gas

pressure and temperature at the inlet to the stage m3/s



6.9 Choice of Number of Cylinders Manufacturers have a variety of cylinder arrangements; for a preliminary assessment it is not worthwhile to do more than assume a double-acting cylinder and ignore the piston rod effect on the grounds that it is negligible for LP cylinders and that HP cylinders are fitted with tail rods. Estimate the number of cylinders based upon the following considerations: (a) An even number of cylinders is preferred for horizontal-opposed

configurations.

(b) Complete separation of services is required on multi-service compressors. (c) The cylinder size is limited by the distance between crankshaft throws. As

a first estimate take the maximum permissible piston diameter as 25, where 5 is the stroke. For single cylinders, or for two cylinders arranged in 'V', 'L' or horizontal-opposed configurations, take the limiting piston diameter as 35.

(d) High capacity compressors normally have cylinders operating in parallel:

do not select more than four cylinders per stage. (e) The piston rod load is limited. The convention is to ignore inertia load and

consider only the load due to the differential gas pressure across the piston. This load is the chief design parameter determining motion work size: values chosen by a manufacturer for his range of standard machines fix his possible arrangements of cylinders. There is not a unique arrangement for a given process duty. Compressors with rod loads exceeding 800 kN are not commercially available; as a rough guide take the allowable load as:

1.50 S2.5 x 10-4 kN

where 5 is the stroke in mm. The piston rod loads should be equalized, so that the ratio of the largest to the smallest load does not exceed 2.



7 DRIVER TYPE 7.1 Electric Motors For ratings above 1 MW, the preferred driver is a direct-coupled synchronous electric motor with a single outboard journal bearing. For ratings below 1 MW the preferred driver is a direct-coupled induction motor having two bearings. For ratings of 220 kW and below specify a belt drive and the motor speed as the next lower standard less than: 9,500 2/3 rev/s E Where E is the motor rating in kw 7.2 Steam Turbines Geared steam turbines may be used but are not recommended for powers less than 1.5 MW; their use for powers less than 600 kW is restricted to non-condensing operation. 7.3 Special Drivers The following drivers require investigation outside the scope of this Design Guide: (a) Geared gas turbines (b) Direct-coupled or geared Diesel or dual-fuel engines (c) Integrated engine/compressor units, where Diesel or dual-fuel engine

cylinders and compressor cylinders share a common crankcase and crankshaft

(d) Hydraulic turbines



8 VESSELS The inter-stage vessels and pipework are arranged by the machine vendor. However, the gas supply to the first stage is usually taken from a catchpot which is customarily regarded as the capacity element used to attenuate gas pressure pulsation. For this purpose: (a) The effective capacity of the vessel shall exceed 30 swept volumes of the

stage 1 cylinder, or of any stage 1 cylinder where there is more than one.

(b) The length of the connecting pipe between the vessel and a cylinder (assumed double-acting) shall not exceed:

C m 25 x N where C is the velocity of sound in the process gas at

inlet conditions m/s

N is the crankshaft speed rls If these conditions cannot be fulfilled then specify the provision of a separate inlet pulsation damper by the machine vendor. Wire-mesh demisters are not acceptable for such catchpots unless special measures are taken to catch loose wire segments before the cylinder inlet.



APPENDIX A RELIABILITY CLASSIFICATION (abstracted from GBHE-EDG-MAC-5100) Installations having high availability are classified as follows: Class I A Class I installation achieves high availability by having units of high intrinsic reliability, and is characterized by:- (a) The machine being a single unspared unit upon which the process stream

is wholly dependent. (b) The plant section having a single process stream with a long process

recovery time after a shutdown so that the loss of product owing to a machine stoppage is large even though the shutdown is for a short time.

(c) A capability of continuous operation within given process performance

tolerances over a period of more than three years, without enforced halts for inspection or adjustment.

(d) Component life expectancies exceeding 100,000 hours operation. Class 2 As for Class I but where infrequent plant shutdowns of short duration are acceptable because the process recovery time is short. Consequently the period of continuous operation capability can be reduced and is taken as 4,000 hours for this classification. Class 3 As for Class 1 or 2 but with a machine performance deterioration during the operating period accepted, or countered by adjustment of process conditions or by other action on the part of the plant operators.



Class 4 A Class 4 installation achieves high availability by redundancy and is characterized by having: (a) One or more machines operating with one or more standby machines at

instant readiness at all times to take over automatically upon malfunction of a running machine.

(b) Operating and standby machines designed specifically for their functions

so that they are not necessarily identical. (c) Component life expectancies exceeding 25,000 hours operation. Class 5 A Class 5 installation follows the Class 4 redundancy concept and is characterized by: (a) One or more machines operating with one or more identical machines

installed as spares to take over the process duty at the discretion of the plant operators.

(b) One or more machines operating in plant sections where product

storage is sufficient to give the plant operators adequate time to assess the malfunction and take remedial action. Alternatively, plant sections where a single machine stoppage does not cause a disproportionately large process upset.

Class 6 Machines intended for batch or intermittent duty. Where high demand reliability is essential, the machines lie in Class 4.



APPENDIX B CONDITIONS FOR LUBRICATED CYLINDERS & GLANDS B.I CHOICE OF OIL VISCOSITY GRADE FOR CYLINDERS Note that the oil film on the cylinder wall is very thin 9~ 10 µm). The procedure is as follows: (a) Estimate Oil Film Temperature

For typical machines having water-cooled jackets, take: T 0 = 2 Ti + Td for double-acting cylinders

3 Or T 0 = 5 Ti + Td for single-acting cylinders 6 Where To = oil film representative temperature 0C Ti = inlet gas temperature Td = discharge gas temperature (b) Estimate Oil Film Pressure The representative pressure (Po) in the oil film is taken as the mean effective pressure for double-acting cylinders. Po = Pi 1 + n r(n-1)/n - 1 n - 1



where Pi = inlet gas pressure bar abs r = compression ratio across the cylinder n = compression index, taken as the ratio of specific heats Cp/Cv Single-acting cylinders have a balance chamber: two single-acting cylinders may be combined in tandem with a pressured balance chamber between them. For such configurations take Po as the arithmetic average of the balance chamber pressure and the mean effective pressure calculated as above. (c) Find Dissolved Gas Content

Gas solubility is defined by the Ostwald coefficient (the volume of gas dissolved in unit volume of liquid, both volumes being measured at the same temperature and pressure). The solubility is substantially independent of oil type and viscosity.

Ostwald coefficients for some common gases are shown in Figure B.1



The gas/oil volume ratio (λ) is given by: λ = Ostwald coefficient x Po 293 To + 273

(c) Estimate Effective Oil Viscosity

Ѵ o = A 1 + Po Ѵ b 340 Where Ѵ o = effective oil viscosity cSt Ѵ b

= oil viscosity at 1 bar abs and To cSt for one of the oil grades shown on the ASTM chart (remember that this chart uses Fahrenheit temperatures). The constant A and the index b depend upon i\ and are given in Fig B.2.



(e) Choose Viscosity Grade Figure B.3 charts the limiting viscosity grade for the maximum operating pressure drop across the piston.



FIGURE B.3 OIL VISCOSITY Operation in Zone A is unrestricted. Operation in Zone B is permitted only when carbon-filled PTFE rings are used in conjunction with the oil lubricant. Operation in Zone C is forbidden. B.2 CHOICE OF VISCOSITY GRADE FOR PACKINGS Current practice is to use the same method as that for cylinders. For the common segmental packing arrangement, the oil film on the rod travels into the low pressure zone and eventually into the atmosphere. Treat as a single-acting piston where: T I = gas temperature at cylinder inlet or in the balance chamber °C T d = maximum ambient air temperature, taken as 32°C for UK installations °C P o = geometric average of cylinder discharge pressure (or maximum balance

chamber pressure) and the packing discharge pressure bar abs Specify the same grade of oil for lubricating both cylinder and packing.



Figure B.4



APPENDIX C ESTIMATE OF LUBE OIL CONTAMINATION OF PROCESS GAS C.I Oil Injection Rate Provided that the oil viscosity grade is properly chosen it is only necessary to inject sufficient oil to keep the cylinder and packing surfaces coated. The actual feedrate is then given by

Ø x D x U grams/hour Where D is the diameter of the piston or rod mm

U is the mean piston speed m/s Ø is the specific mass feed rate from Fig C.I



These feed rate values should be increased to cater for the oil removed as vapor by some process gases, e.g. natural gas, which deplete the oil film. C.2 Oil Carry-Forward Rate C.2.1 Total Rate For estimating contamination of the process gas take the total input as the sum of the oil injected into the cylinder and 20% of the oil injected in the packing. Oil injected into stages preceding the last stage can be neglected provided that there is a condensing intercooler and separator before the last stage. C.2.2 Condit ion of Oil Oil is carried forward in two forms: (a) as droplets produced by shear at the valves, having a predominant

particle size about 5µm. These cannot be satisfactorily removed by conventional separators.

(b) as vapor content which can be reduced only by cooling.

The vapor content is estimated by an empirical method, using the saturation vapor phase enhancement factor f given by:

Loge f = 0.012 P Mo + 10000 x H ρ0 Tc where

P = gas discharge pressure bar abs T = absolute temperature of discharge gas K M o = mean molecular weight of the oil Ρ o = density of oil at temperature T kg/liter



H = enthalpy difference of the gas at temperature T kJ/kg-mol between pressure P and 1 bar

Tc = critical temperature of gas K For a mixtures of gases, calculate the enhancement factor for each component and obtain f for the mixture as:

F = 1 + (f1 - 1) + (f2 -n 1) + …. Then p* = f x p Where: P is the oil vapor pressure in air at bar abs

atmospheric pressure and at temperature T (K) p* is the oil vapor pressure in the gas bar abs



APPENDIX D INFLUENCE OF GAS COMPOSITION AND MACHINE CONSTRUCTION ON FILLED-PTFE PISTON RING SEALS The current state of knowledge is empirical and can be summarized as follows: D.1 Influence of gas composition (a) Experience in ordinary air compressor service is not valid for other gases.

The wear rate for carbon-filled PTFE rings in oxygen and air is higher than in inert gases (nitrogen, helium) or in quasi-inert gases {carbon dioxide, chlorinated hydrocarbon refrigerants).The wear rate in methane and ethylene is especially sensitive to the composition of the ring material.

(b) The wear rate increases below a critical content of water vapor

corresponding roughly to a dewpoint of -40°C. D.2 Influence of machine construction (a) The wear rate of dry PTFE depends on the formation of a transferred film

on the counter surface. If this is prevented the wear rate can be catastrophic. Consequently:

- the combination of an unlubricated cylinder with a lubricated gland

is forbidden. - condensation within the cylinders of unlubricated machines shall be

avoided. This requirement may lead to the provision of a separate jacket cooling system with temperature control to avoid over-cooling.

(b) Experience suggests that longer lives are obtained from piston rings than

from rider pads. Consequently vertical cylinders are preferred.



APPENDIX E LIMITS ON GAS TEMPERATURES E.I Lubricated Cylinders These limits are based on formation of carbonaceous deposits when using conventional straight mineral oil to BS 4475 CSB.

The use of synthetic hydrocarbon or esters as lubricants permits higher temperatures but each application needs special investigation. These temperature limits are appropriate to normal operation; they may be exceeded for short periods, eg upon a valve fault before the machine is shut-down for valve replacement. E.2 Limiting Temperatures for Carbon-Filled PTFE Rings The criteria for life expectancy are the mean temperature of the piston ring and the differential pressure across the piston. By convention, the mean temperature (T p) of the piston ring is taken as the arithmetic average of the inlet and discharge gas temperatures, for water cooled cylinders. When ΔP< 14 bar, Tp < 105°C When ΔP> 14 bar, Tp <. [277 - 150 Log10 ΔP] o C



Even though the mean temperature of the ring is acceptable, there is a secondary effect of the gas discharge temperature. For gas discharge temperatures between 160 o C and 200 o C the life expectancy of PTFE-based rings will be reduced. Above 200 o C specify carbon rings. Process gases which attach carbon require investigation beyond the scope of this Design Guide. If the PTFE rings are oil-lubricated the temperature limit changes to: T p < [546 - 288 log10 ΔP] or T p < 120 o C whichever is the lower Compressor Process Sketch



DOCUMENTS REFERRED TO IN THIS ENGINEERING DESIGN GUIDE This Engineering Design Guide makes reference to the following documents. BRITISH STANDARDS BS 4475 Straight Mineral Lubricating Oils (referred to in Clause E.1) AMERICAN PETROLEUM INSTITUTE API 618 Reciprocating Compressors for General Refinery Services

(referred to in Figure 3). ENGINEERING DESIGN GUIDE GBHE-EDG-MAC-5100 Reliability Analysis - the Weibull Method (referred to

in Appendix A).

