multhiphase pumps for minas losf.pdf

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Copyright 1999, Society of Petroleum Engineers Inc.

This paper was prepared for presentation at the 1999 SPE Asia Pacific Oil and GasConference and Exhibition held in Jakarta, Indonesia, 20–22 April 1999.

This paper was selected for presentation by an SPE Program Committee following review ofinformation contained in an abstract submitted by the author(s). Contents of the paper, aspresented, have not been reviewed by the Society of Petroleum Engineers and are subject tocorrection by the author(s). The material, as presented, does not necessarily reflect anyposition of the Society of Petroleum Engineers, its officers, or members. Papers presented atSPE meetings are subject to publication review by Editorial Committees of the Society ofPetroleum Engineers. Electronic reproduction, distribution, or storage of any part of this paperfor commercial purposes without the written consent of the Society of Petroleum Engineers isprohibited. Permission to reproduce in print is restricted to an abstract of not more than 300

words; illustrations may not be copied. The abstract must contain conspicuousacknowledgment of where and by whom the paper was presented. Write Librarian, SPE, P.O.Box 833836, Richardson, TX 75083-3836, U.S.A., fax 01-972-952-9435.

AbstractMultiphase pumps are modified liquid pumps that are capable

of pumping various combinations of oil, water, gas, and sand

in the same pipeline without separation. Minas Light Oil

Steam Flood (LOSF) Pilot Project will utilize the multiphase

technology to boost casing fluid with high steam and gas

content from wells to the Gathering Station without prior

separation. Hence, tremendous savings in capital and

operating costs are realized from strategic placement of

multiphase pumps since large separators, liquid pumps and

compressors are eliminated.Two uniquely different multiphase pumps, a 500 hp twin-

screw pump and a 700 hp helico-axial pump, were recently

tested and qualified at the Texaco Multiphase Flow Facility,

where flow rates as high as 145,000 BPD and gas volume

fractions (GVF) up to 100% were achieved. An illustration of

each pumping unit is given followed by a presentation of

measured performance data relative to the operating conditions

chosen by Caltex Pacific Indonesia (CPI). Based on test

results, the advantages and disadvantages of the competing

technologies are compared, noting that CPI will eventually

consider these data to choose either twin-screw or helico-axial

technology for the Minas Light Oil Steam Flood project.

Results obtained at the Texaco Multiphase Flow Facilityindicate that twin-screw pumps are 25% more efficient than

helico-axial pumps. Furthermore, twin-screw pumps are

virtually insensitive to liquid slugs and large changes in inlet

gas density; helico-axial pumps require slug catchers upstream

of the pump inlet to absorb these intermittent slugs and large

variations in GVF. It is anticipated, however, that helico-axial

pumps are superior in resisting long-term wear from

continuous exposure of sand, which will be investigated in

Duri in the fourth quarter of 1998.

IntroductionThe nature of steam flood in Minas field is quite unique since

the light Minas oil (32° API) is so volatile that casing vapors

must be produced. Production streams from both casing and

tubing must be piped to the Bulk Metering Station and thenusing the existing lines, to the gathering station. In these lines

LOSF flow will be quenched by ‘colder’ fluid from other parts

of the Minas field which are water injected.

Simulations show a significant percentage of the oi

production will be vapors. Closed casings would severely

hamper production rates, and vapors would be produced

through the tubing anyway. Having vapors in tubing is very

undesirable since this will seriously affect the performance of

multi-stages centrifugal pump downhole. Therefore a casing

vapor collection is required.

In order to promote both high rates of inflow from the

reservoir and natural flow, Minas wellhead casing pressures

must be drawn to a relatively low value, 60 psig. Howeverpreliminary calculations showed that the existing gathering

system would not support the expected amount of these steam-

dominant vapors at pressure below 60 psig. Some alternatives

could be adopted to solve this problem, among them are:

• Cool the casing vapors at locations near the well heads.

• Use multiphase pumps to boost casing streams to

gathering station.

The first option means that HC condensates and water mus

be separated from steam. This choice would requires various

different equipment, higher OPEX (energy cost especially) and

more intensive maintenance efforts. If a separate line would

not be built to transport the liquids to the gathering station, the

stream must be piped to the existing lines. Hence, this coolingwould be inefficient since the streams will be quenched

anyway. Therefore, LOSF design team decided to implemen

the later alternative.

Multiphase pump application becomes critical when steam

including light oil vapors reach producing wells. This will

happen 1.0 to 2.5 years after injection. The implementation o

multiphase pump is expected to:

SPE 54294

Multiphase Pumps for Minas Light Oil Steam Flood, Sumatra-IndonesiaErwinsyah Putra, Caltex Pacific Indonesia and Donald D. Uphold, Texaco

http://contents.pdf/



2 E. PUTRA, D.D. UPHOLD SPE 54294

• Boost casing stream to tubing stream pressure

• Reduce volumetric flow rate of casing fluids through

condensations across the pump

• Reduce wellhead pressures and therefore enhance well

productions.

The Minas multi phase pump must also be capable of

handling:

• Unsteady flow rates and terrain slugs• Sand production

• High temperature (up to 320 °F)

Multiphase Pump SelectionTwo commercially available technologies are being

considered, twin-screw and helico-axial.

Details of each technology will be presented later. To

determine which technology is most suitable for Minas-LOSF

application, each pump would undergo several tests to evaluate

its performances. The design point is described as follows:

• Suction pressure: 75 psig

• Differential pressure: 100 psi

• Gas Volume Fraction/GVF: 75%

• Capacity: 125,000 ABPD

The first test was conducted at Texaco Flow Loop Test

Facilities at Humble. The test matrix consisted of:

• Design point test

• High speed test

• High GVF test (up to 100% GVF, run dry)

At Humble, the process fluid was homogeneous multiphase

mixture of Fuji crude oil (30° API), water and methane heated

to 200 °F.

The final test will be conducted in Duri field in last quarter

of 1998. Both pumps will be exposed to multiphase mixtures

containing oil, water, gases (including CO2 and H2S), wetsteam and sand. The temperature may reach as high as 300 °F.

Data from Humble and Duri plus pump’s economics will

determine which technology is best for Minas-LOSF Pilot

Project.

Helico-axial TechnologyA 7-stage rotodynamic pump equipped with a 700 hp motor

and a Variable Frequency Driver was chosen for testing.

Figure 1 and 2 illustrate the helico-axial pump. This pump is a

horizontal seven-stage multiphase pump of the helico-axial

design. The collection of stages accommodates variable

reduction in gas volume flow rate as the fluid moves through

the pump. The hydraulic components are designed to preventgas-liquid separation, which would otherwise produce

diminished head, or vapor lock, as is common in conventional

centrifugal pumps.

Each stage of the pump consists of a rotating helico-shaped

impeller and a stationary diffuser, yielding a hybrid between a

dynamic pump and an axial compressor that allows a wide

range of liquid flow rates and inlet gas concentration.

Each impeller provides pressure boost and the interstage

diffuser homogenizes and redirects the flow for the next set of

impellers. This interstage mixing increases the overal

efficiency and enables stable pressure-flow characteristics.

The nature of open-axial hydraulics allows higher tota

flow rates compared to their competitive counterpart twin

screw pump. This is especially important when considering

the production life of the field for maximizing oil recoveryAdditionally, open-axial passages accommodate solid particle

and minimize accumulation in the pump casing. Finally, open

axial flow minimizes the risk of mechanical seizing due to

overheating at high GVF.

In theory, a rotodynamic pump creates pressure

dynamically, where shaft torque is converted into angular

momentum. Hence, differential pressure depends on moto

speed and inlet fluid density. This makes rotodynamic pump

extremely sensitive to small changes in inlet conditions

Hence, large changes in shaft torque under intermittent flow

are common with rotodynamic pumps. Three features are

offered to smooth these disturbances: mechanical design

improvements, flow homogenizer (buffer tank), and variablespeed capabilities. Firstly, the impellers are stacked with a

small interference shrink fit in order to maintain a smal

residual unbalance. The torque is transmitted to the impeller

with a single key or pin. Two (2) anti friction roller bearings

stabilize radial thrust and one tilting-pad thrust bearing carries

the axial load. Secondly, the buffer tank, or static mixer, i

installed upstream of the pump to accommodate slug flow

The buffer tank absorbs energy of liquid slug fronts and

smoothens out fluctuations in gas density and pressure. Hence

repetitive torque changes within the pump are minimized

Finally, the variable speed drive (VFD) adjusts the motor

speed to maintain constant inlet pressure. Differentia

temperature across the pump casing is expected to increasewith increasing GVF. Special care has been taken to minimize

this differential temperature (i.e., internal design modification

thermal insulation) as it could potentially damage the rotor and

diffusers due to thermal distortion of the shaft.

The rotodynamic pump utilizes only one rotating shaft, or

rotor. The pump rotor (with impellers), diffusers, bearings

shaft seals and cover are designed as a cartridge assembly

which is introduced into the casing to reduce maintenance

downtime. The single shaft principle is simple and permits

higher rotational speeds (up to 3,600 rpm for this specific unit

when compared to the twin-screw counterpart (up to 2300

rpm). Process fluid enters through a vertical inlet. The initia

stage homogenizes the fluid and prepares it for the firsboosting stage. Fluid is then pumped axially via seven

impeller/diffuser stages, where the fluid builds pressure

dynamically. The hydraulic stages are sized according to the

principle that most of the compression occurs during the later

stages. Process fluid then discharges vertically. Recall that a

tilt-pad thrust bearing assembly is installed at the non-drive

end of the pump to handle the axial load and maximize roller




consumed to compress Duri gas, developing less differential

pressure at a chosen speed on the diagram. These effects will

be further studied in Duri.

The pump system performance at this high ∆P condition is

quite impressive. It appears the system was intended to

operate at this point rather than the design point. Furthermore

as illustrated in Fig. 5, efficiency is optimized when operating

near this condition, suggesting the BEP occurs near the high-∆P point. Alternatively, this finding confirms a safety factor to

withstand routine wear from elevated temperatures and solids

content while operating at the design point for an extended

period of time. If this is the case, this pump should fare much

better than the twin-screw pump over the production life in the

Minas LOSF.

The pump successfully demonstrated operation up to 94%

GVF with very little discharge temperature build-up over the

suction temperature. However, the pump was only capable of

developing 25 psi differential pressure at this GVF. In

addition, hydraulic efficiency dramatically dropped to 16%.

Attempting to increase the GVF above 94% caused inlet

pressure to drop below 60 psig, tripping the pump on lowsuction pressure. These results question the effectiveness of

rotodynamic pumps at high GVF. Buffer tanks located

upstream of the pump may smooth off intermittent gas spikes

(as long as liquid is initially present in the tank).

Figures 6 and 7 contain families of curves generated from

test data. Figure 6 is differential pressure ∆P (psi) versus total

flow rate; Qtot (bpd) at various pump speeds (rpm). Inlet

pressure is 75 psig and GVF is 75 %. Note that the slopes of

the curves increase with increasing rpm. This confirms that

rotodynamic behavior approaches pure axial-compression as

motor speed becomes large. At lower speeds, the flat curves

depict a hybrid between centrifugal and axial hydraulics. This

suggests low-rpm curves should exhibit inflection points upondecreasing flow rate (this requires measurements at flow rates

below those in Figure 4 for verification).

Figure 7 illustrates ∆P versus Qtot while varying GVF at

constant speed of 3575 rpm. Note that ∆P dramatically

diminishes upon increasing GVF. As with decreasing rpm

above, increasing GVF here causes more hydraulic slip upon

increasing ∆P. Our string tests conclude that maximum GVF

(without a buffer tank) is 94%. A buffer tank, or other liquid

source, preserves a reasonable operating range by keeping the

GVF below 90%. In addition, VFD maintains desired inlet

GVF conditions.

Figure 5 is ηmultiphase versus ∆P. Inlet pressure is 75 psig

and GVF is 75%. Measured data indicate BEP (37%) occurs

near 140 psi and 3350 rpm not at the design ∆P (30%). This

arises from additional safety margin incorporated in the pump

design, allowing the pump to withstand long-time wear from

aggressive sand erosion and maintain intended design

condition during extended use. These pumps inherently have

large clearances between rotating and stationary parts, thus

maximizing sand throughput and minimizing erosion (and

consequently minimizing chances of mechanical seizure from

large thermal growth during high GVF operation). These

characteristics of long-term endurance highlight the strength

of rotodynamic pumps over twin-screw pumps. These features

will be verified in the Duri later this year.

Twin-Screw Pump Test Results

The pump performed exceptionally with the test requirementsof the purchase order and the test matrix, table 2. This

includes process and mechanical requirements. At design

speed and moderate differential pressure (up to 150 psi)

measured and theoretical flow rates are almost equal, Fig. 8

Hence, twin-screw pumps exhibit high volumetric efficiency a

conditions near the design point, Figure 9.

The pump system was successfully operated continuously

for 3 hours at maximum design speed and 75 psi differential

pressure. In particular, pump vibrations were measured at 0.13

ips, well below the required limit of 0.3 ips. Motor-winding

and pump-bearing temperatures asymptotically approached

150 F and 108 F, respectively, after only two hours of

continuous operation. The manufacturer suggested reducingthe pump’s differential pressure from 100 psi to 75 psi to keep

the motor power below the design power rating. The

specification requires manufacturer to guarantee design

conditions only, which equates to about 367 hp. These result

are promising for occasional long-term operation at high load

conditions. It is not expected in the Minas LOSF to frequently

run at these conditions, allowing a comfortable safety margin

for long run times.

The test proved that the pump is capable of operated

continuously for 1 hour at 100 % GVF and 125 psi differential

pressure. Pump vibrations were measured at 0.08 ips. Fluid

temperature rose from 175 F inlet to 220 F outlet, a 45 F

increase. These results impressively demonstrate the pump’scapabilities at high GVF continuous operation. The interna

liquid recirculation (screw sealant) allows a safety cushion of

approximately one-hour of continuous operation during

intermittent periods of dry gas. In real case, it is anticipated

that steam condensation will increase this safety margin.

Figure 8 is total flow rate Qtot (bpd) versus differentia

pressure ∆P (psi). Measured data in the range of 100 - 150 ps

differential pressure correspond to 115,000 – 130,000 bpd

which matches quite well with the pump curve supplied by

manufacturer. The variation in measurement is caused by

differences in GVF. At the same differential pressure, flow

rates at 80% GVF are larger than observations at 70% GVF

Note that as differential pressure increases from 150 – 230 psitotal flow rate drops to 95,000 bpd. A fraction of this decrease

is due to slip, however the majority of it arises from the wide

range of motor speeds used to generate this curve. In general

lower flow rates correspond to lower motor speeds. The effec

of slip is more apparent in Figure 10.

Figure 10 is total flow rate Qtot (bpd) versus motor speed

(rpm). A conceptual linear relationship is plotted against the



SPE 54294 MULTIPHASE PUMPS FOR MINAS LIGHT OIL STEAM FLOOD, SUMATRA-INDONESIA 5

measurements done in the range 1500 – 1950 rpm. The

variation in measurement is a reflection of slip since the data

are taken over a wide range of differential pressures and motor

speeds. At 1800 rpm, total flow rate varies from 105,000 bpd

to 115,000 bpd with differential pressures ranging from 100 –

150 psi. Moreover at 1500 rpm and 100 psi, total flow rate

varies from 90,000 bpd to 95,000 bpd. Slip is calculated as

follows:• At 1800 rpm and 150 psi, percent slip is

• At 1500 rpm and 100 psi, percent slip is

With a higher ∆P and slip at 1900 rpm, these calculations

suggest that slip is more influenced by changes in differentialpressure than changes in motor speed. Regardless, at constant

motor speed, slip increases with increasing differential

pressure; at constant differential pressure, slip increases with

decreasing motor speed.

Figure 9 displays total multiphase efficiency versus

differential pressure at constant inlet pressure = 75 psig and

constant speeds of 1500, 1900, and 2100 rpm. Maximum

efficiency or BEP is 44% and occurs near the design point at

1900 – 2100 rpm. The efficiency curve for 1500 rpm is

approximately 10% lower than the curve at 1900 rpm. Hence,

for speeds at or below the design speed, a 20% decrease in

motor speed corresponds roughly to a 10% decrease in pump

efficiency. For speeds above the design speed, such as 2100rpm, the optimum efficiency occurs at a value slightly less than

the design efficiency. Note that efficiency at 2100 rpm

decreases as 125 psi is reached. Also note that efficiency

increases with increasing rpm at ∆P slightly lower than design

∆P.

ConclusionsThe following table summarizes test key-results:

Twin-screw Helico-axial

Efficiency at

Design point

44% 30%

Shaft power atdesign point

325 hp 500 hp

Maximum GVF 100% 94%

Test data also shows, for twin-screw pumps, efficiency slightly

decreases with increasing GVF. Meanwhile for rotodynamic

pumps, efficiency drop-off is more dramatic.

A head-to-head comparison between the twin-screw and

rotodynamic pump based strictly on results obtained a

Humble is inconclusive to determine which is the bette

technology. The twin-screw pump is hydraulically more

efficient than the rotodynamic pump (44% versus 30%)

therefore OPEX is lower. The twin-screw pump is also

capable of handling slugs and 100% GVF without the need for

buffer tanks located upstream of the pump. In addition to thatinternal seal flush and lubrication system of the helico-axia

pump is more complicated and requires careful maintenance.

However, rotodynamic pumps are suppose to be superior

in high sand environments that suffer rapid temperature

changes due to process upsets and steam condensation. The

internal clearances between rotating impellers and stationary

parts in rotodynamic pumps are much larger than clearances in

screw pumps, enhance throughput of sand and relax therma

deflection criteria. These effects should be related to better

mechanical seal reliability and an extended bearing life

Consequently OPEX is lower since the pump requires less

maintenance and downtime. Although we foresee difficulty

with the pump in severe sand environments, short-termeconomics will probably favor this technology over

rotodynamics based entirely on CAPEX.

Each pump will undergo a thirty-day field test in Duri

Both pump will be exposed to multiphase mixtures at elevated

temperature (up to 300 °F). This multiphase mixtures contain

production fluid (80% WC), wet steam (70% quality) and sand

(up to 0.2% by liquid volume). Duri test-results should yield

additional important-data to declare best available technology

for Minas Light Oil Steam flood project.

AcknowledgmentsThe following individuals gave instrumental contributions in

making this project completed, Lee Larson, John McSharryfrom Chevron, Joe Anderson from Texaco and Kevin Kassner

Hadi Prijono from Caltex Pacific Indonesia.

References1. Uphold, D. D. 1998 Multiphase Twin-Screw Pump

Performance Test, Texaco EPTD, July, Houston, TX, Repor

#98-0107

2. Uphold, D. D. 1998 Multiphase Rotodynamic Pump

Performance Test, Texaco EPTD, September, Houston, TX

Report #98-0151

%9%100000,115

000,105000,115 =

−= x slip

%5%100000,95

000,90000,95 =

−= x slip




Appendix : Multiphase Efficiency

It is worthwhile to investigate multiphase pumping efficiencies to establish estimates for baseline comparisons with other hydraulic

technologies, such as gas-lift, ESPs, compact separators, surface liquid pumps and gas compressors. In general, multiphase pumping

efficiencies are 30 – 50 %, liquid pumping is 50 – 70 %, gas compression is 70 – 90 %, and gas-liquid separation is above 90 %

Project economics will suggest if savings derived from conventional component efficiencies will offset their high CAPEX cost.

Multiphase pumps are less expensive than conventional technologies, yet operating cycles and current MTBF of multiphase pumps arenot yet appealing to production operators. These issues define our goals to improve and deploy multiphase pumps throughout the

industry, or search for more appealing alternatives.

˚ In this section, we introduce multiphase efficiency (ηmultiphase) in terms of actual shaft horsepower (hpshaft ) and multiphase hydraulic

horsepower (hpmultiphase). Multiphase efficiency is a single parameter that accounts for the individual contributions from liquid pumping

and gas compression (polytropic):

where z is compressibility factor, Qgas,std is gas flow rate (mmscfd), Qoil is oil flow rate (bpd), Rs is gas solubility (scf/st.bbl), T s is

suction temperature (F), Ps is suction pressure (psig), Pd is discharge pressure (psig), and γ is the ratio of gas specific heats. Note tha

the choice of polytropic compression is arbitrary. The polytropic assumption is acceptable for large GVF. If GVF is small, then the

thermodynamic path resembles isothermal compression. Since we are interested in GVF > 75%, the polytropic process is more

appropriate than the isothermal process. hpliquid and hpshaft can be calculated as follows:

hpliquid is calculated from the standard hydraulic pump equation modified for two-phase flow, where Qwater and Ooil are water and oi

flow rates (bpd), ∆P is differential (psi), and Bo is the formation volume factor of oil (bbl/st.bbl). Gas compression follows a

polytropic process: Power consumed by the motor is determined from measured and known electrical properties, where V is volts, A is

amps, p f is power factor and hpshaft is motor efficiency.

Table 1 : Helico-axial Pump Test-Results

Test Time, hr Qtot, bpd Speed, rpm ∆P, psi GVF, % hpshaft, hp ηmultiphase , %

Design point 1 125,000 3050 100 75 520 30

High-speed 3 145,000 3575 75 75 700 20

High-∆P 1 125,000 3245 140 75 570 37

High-GVF 1 125,000 3550 25 94 250 16

Table 2: Twin-Screw Pump Test ResultsTest Time, hr Qtot, bpd Speed, rpm ∆P, psi GVF, % hpshaft, hp ηmultiphase, %

Design point 1 125,000 1925 100 75 330 44

High-speed 3 145,000 2280 75 75 390 42

High-GVF 1 125,000 1925 100 100 360 39

ηmultiphasemultiphase

shaft

hp

hp=

( )( )P B Q P Q hp o oil water liquid ∆+∆= 000017.0 ( )( )( )( )hp

V A pshaft

f motor

=3

746

η

liquid gas multiphase hp hp hp +=

( ) ( )

−+

+

−

+

−=

−

17.14

7.14

1460

100857.0

1

6,

s

d s

oil s std gas gas

P

P T

Q R Q z hp

γ

γ

γ

γ




_____________________________________________________________________________________________

Figure 1 : Helico-axial Pump Diagrams

Figure 2 : Details of a Helico-axial Pump Stage

Rotating flow impeller

Stationary diffuser



8 E. PUTRA, D.D. UPHOLD SPE 54294 _________________________________________________________________________________________________________

****

* *****

****

*

**

*** **

****

** *

****

High-pressure test

Design-point test

High-speed test

Figure 3: Twin-screw Pump Diagrams

Figure 4 : Pump Manufacturer’s Curve and Test-Points Loci - Helico-Axial Pump




_____________________________________________________________________________________________

Figure 5 : Differential Pressure vs. Multiphase Efficiencyat Various Speeds - Helico-Axial Pump

Figure 6 : Total Flow Rate vs. Differential Pressureat Various RPM - Helico-Axial Pump

Figure 7 : Total Flow Rate vs. Differential Pressureat various GVF – Helico-axial Pump



10 E. PUTRA, D.D. UPHOLD SPE 54294 _________________________________________________________________________________________________________

Figure 8 : Differential Pressure vs. Total Flow Rate – Twin-Screw Pump

Figure 9 : Differential Pressure vs. Multiphase Efficiency – Twin-Screw Pump

Figure 10 : Pump Speed vs. Total Flow Rate – Twin Screw Pump

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