Integrating Special Purpose Compressors

GBH Enterprises, Ltd.

Engineering Design Guide:
GBHE-EDG-MAC-1033

Integration of Special Purpose
Reciprocating Compressors into a
Process

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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
2.2

Preliminary Choice of Machine Type

2.3

3

Parameters

Review of Other Types of Compressor

CHOICE OF NUMBER OF COMPRESSORS
3.1

Driver Considerations

3.3

4

Influence of Reliability Classification

3.2

3

Deterioration of Standby Machines

EFFECTS OF PROCESS GAS COMPOSITION
4.1

Particulate Contamination

4.2

Droplets in Suspension

4.3

Polymer Deposit

4.4

4

Molecular Weight Variation


4.5
4.6

Gas Dryness

4.7

5

Compressibility Variation

Gas Solution in Lubricating Oil for Cylinder and Gland

THROUGHPUT REGULATION
5.1

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

6

Inlet Line Throttle Valve

5.2

5

Hybrid Regulation Systems

PRINCIPAL FEATURES
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

6

Choice of Number of Cylinders


7

DRIVER TYPE

7

7.1
7.2

Steam Turbines

7.3

8

Electric Motors

Special Drivers

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
100 x M x Pi

m3/s

W

- Mass flowrate

kg/s

Ti

- Inlet gas absolute temperature

K

Pi

- Inlet gas absolute pressure

bar abs

Zi

- Gas compressibility corresponding to Pi , Ti

R

- Gas constant of 8.3143

M

- Gas molecular weight

kJ /kg-mol K

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 =
Where

(c)

Zi x R x Ti Loge Pd
M
Pi
Pd = maximum discharge pressure bar abs

Estimate the power required as:

E = 1.6 W x Ht

2.2

kJ/kg

kW

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)

(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)

(c)

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

The maximum compression ratio of any stage is 4.5

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

(2)
(d)

The maximum differential pressure is 1.4 bar
The inlet actual volume flow is within the range 0.1 to 1.0 m3 /s

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 cutoff 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:

m3

60 x Q
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
dP is the chosen differential pressure
band for the receiver

5.3

bar abs

bar

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 Vn =

m-I

TxP

m

constant

=

constant

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 =

where

(1 + α)ᴕ
1 + α xᴕ

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
1/2

- 0.4 (1 - ᵑv)

DxN

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
ρ½

Pi (r – 1)
r

3/8

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
Where P is the differential pressure across the cylinder

m/s
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
2&4
5&6

Mean Piston Speed
m/s
4.5
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
across the piston

bar

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
Ps - Po

Where

> 0.3

ΔP s is the variation in pressure

bar

Within the seal box

Ps

is the nominal pressure
within the sealbox during the
piston rod instroke

bar abs

Po

is the nominally constant
pressure at the seal exhaust vent.

bar abs


(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
32 x δ

0.84 – 0.16 x r 1/n*

Where

ᶯv

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

ᶯ

For the first estimate of v 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
ΔP x Q i

½

r/s


Where
Qi

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
U

r/s

- For carbon piston rings

37
U

r/s

6.8

Calculation of Piston Diameter

For a single cylinder, double-acting,

D

=

20 x V I
ᶯv x π x S x N

½

x 104

mm

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
E

2/3

rev/s

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 noncondensing 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
25 x N
where

m

C
is the velocity of sound in the process gas at
inlet conditions

m/s

N

rls

is the crankshaft speed

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
n-1

r(n-1)/n

- 1


where

Pi

=

inlet gas pressure

r

=

compression ratio across the cylinder

n

=

compression index, taken as the ratio of specific heats Cp/Cv

bar abs

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
Ѵ b = oil viscosity at 1 bar abs and To

cSt
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
ØxDxU
Where

grams/hour

D is the diameter of the piston or rod
U is the mean piston speed
Ø is the specific mass feed rate from Fig C.I

mm
m/s


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
ρ0

+

10000 x H
Tc

where
P

= gas discharge pressure

T

= absolute temperature of discharge gas

bar abs
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
between pressure P and 1 bar

Tc = critical temperature of gas

kJ/kg-mol

K

For a mixtures of gases, calculate the enhancement factor for each component
and obtain f for the mixture as:

F
Then p*

=
=

1 + (f1 - 1) + (f2 -n 1) + ….
fxp

Where:
P

is the oil vapor pressure in air at
atmospheric pressure and at temperature T (K)

bar abs

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:
-

(b)

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 overcooling.

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 shutdown 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).


Integrating Special Purpose Compressors

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