1. Centrifugal Pumps - Forsiden - Universitetet i Oslo · 1. Centrifugal Pumps ... Since pumps and...

MEK 4450 Offshore Technology 09.11.2011 Boosting Options

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1. Centrifugal Pumps Pumps belong to what is often called turbomachinery, where turbo is a Latin word meaning spin or whirl. Turbomachinery comprises pumps (also reciprocating ones which do not spin or whirl), compressors and turbines for both liquids and gases. Other machinery may be included, like torque converters which is a kind of hydraulic clutch used in automatic transmissions.

Centrifugal pumps are by far the most common type of pumps both in the process industry and elsewhere, and they come in a large variety of sizes and design. For most liquids and applications a centrifugal pump is the default choice, i.e. the one that is automatically checked first. However, for many metering applications positive displacement pumps would be the first choice. Poe and Bonall (1999) give a selection guide for low-flow pumps,where they define low-flow as less than 0.5 m3/min.

Notation: Since pumps and compressors act on the total mass according to the laws of mechanics and fluid mechanics, mass (kg) or mass flow (kg/s) is the natural choice, not kmoles as for, say, separation calculations. The letter used in most mechanical engineering literature for power is P, but that is used also for pressure here, and confusion might result in some cases. The letter W is used for work, but we want to distinguish clearly between work and work per time (power). Avoiding too difficult typing is also a factor in the choice. So power is given the letter P here, which is different enough from P. Volume flow is traditionally called Q, confusion with heat or heat flow is not likely.

1.1. Impellers

Impellers are sketched in Fig 8.1. Note that the blades are usually curved backwards, as shown here. Sometimes the fluid leaving the impeller flows into an open volute on its way out of the pump, in other cases there are fixed diffusor blades to guide the fluid.

Figure 1.1 Centrifugal pump impellers with and without fixed diffusor.


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Figure 1.2 Close-up image of rotor with detail of inlet. (Courtesy GE Oil and Gas)

The two main pump groups are positive displacement pumps (fortrengningspumper) and dynamic pumps. The first kind displaces a volume of liquid through mechanical movement of some pump part (typical example: a piston moving). The second kind imparts momentum to the fluid, usually by means of a rotating impeller (løpehjul) with blades or vanes (skovler). These two main groups are also used for compressor classification.

Following White (1994) the pumps are further divided as follows:

Positive displacement pumps

• Reciprocating o Piston or plunger o Diaphragm

• Rotary o Single rotor

- sliding vane

- flexible tube or lining

- screw

- peristaltic

o Multiple rotors - gear

- lobe

- screw

- circumferential piston


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Dynamic pumps

• Rotary o Centrifugal (=radial flow) o Axial flow o Mixed axial and radial flow

• Special o Jet pump or ejector o Electromagnetic pumps o Fluid actuated: Gas lift, hydraulic ram

1.2. Design Equations

The incoming liquid flows to the center of the impeller and gets an increasing velocity and hence kinetic energy as it passes the impeller. The pressure stays largely constant through the impeller. After leaving, the fluid velocity falls to normal pipe flow value, and the kinetic energy is converted to displacement energy (PV work), so the pressure goes up as we want. However, if we take an energy balance across the pump, the fluid velocity is basically the same in and out in the piping, and we do not have to include kinetic energy differences. The only terms we need to include are those shown in Fig 8.2. General equations are derived from the Bernoulli relation for flow with an incompressible fluid.

constant*2

1 2 =++++ wf hhP

hguρ

(1.1)

Figure 1.3. Energy balance for a pump.

As an equation , we get the following simple expression:

P = H tot,2 - H tot,1 = ρQ( H2 – H1) (1.2)

ρQ is the mass flow rate, H is the specific enthalpy, Power is in the units associated with the enthalpy. The equation would obviously be the same for a compressor as the velocities there too are about the same in and out.

Usually we know the flow in (amount, composition, T, P). We have specified a pressure out, since that is our desired quantity, and we want to find the power and the outlet T. Thus, we have one equation and want to determine both P and T2. So we need more information. This is where pumps and compressors part way.


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For pumps we can usually assume the liquid is incompressible, and therefore the temperature does not change much across the pump. The internal energy does not depend much on P when T is constant, so we get, using total values:

P = U2 + P2V2 – U1 – P1V1 = V∆P = Q∆P (1.3)

With V = Q in m3/s (remember, it is always flow rates) and P in kPa the power is given in kW.

The power in eqs (1.2) and (1.3) are for frictionless pumps. So we have to introduce the efficiency, which is always defined as (ideal power/real power) for pumps. The real power then becomes

Preal = η-1 Q ∆P (1.4)

To use eq (1.4) we need to know the efficiency, which we will find in a diagram supplied by the manufacturer. (See later)

The power needed by a pump can also be found from looking at the potential energy of the fluid after the pump, if we consider that the pump is used to lift the liquid. We disregard friction and pump efficiency. We can then set, where ∆h is the elevation change

P = mg∆h = ρQg∆h (1.5)

This height is called head (løftehøyde), and plays a key role in pump and compressor literature. The principle is illustrated in Fig. 1.2:

Figure 1.4. Head produced by a pump

The pressure in the liquid behind the pump caused by this head is given by

∆P = ∆hρg = ∆hρg (1.6)

As seen, we can always express a pressure as a corresponding pressure head and vice versa. Either way, the engineer needs to be careful with units. With h in m, g in m/s2 and density in kg/m3 the pressure unit becomes N/m2 [=] Pa, i.e. a large number. The AES standard for representing these equations is more pedantic and enforces the concept of unit conversion.

∆P = ∆hρg/gc = ∆hρg/gc (1.7)

Where gc is implemented to remind the engineer that there exists an implicit unit conversion

• In SI units gc= 1 N s2/(m kg)

• While in the AES standard gc= 32.174 lbf s2/(ft lbm)

For one-dimensional flow in a pipe we can calculate the interchange between pressure and velocity using the Bernoulli equation (1.1). We now need something similar for the impeller. Since


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we have to describe flow across the rotating impeller, we need to use terms from mechanics that are appropriate for rotation.

1.3. Rotational flow

If a force F acts on a freely rotating arm with radius r it gives rise to torque and rotation. The torque (dreiemoment) is given by

τ = F r = r )( tmudt

d (1.8)

where we have used Newton’s second law that force is the derivative w.r.t. time of the momentum (impuls), and ut is the tangential velocity.

Figure 1.5 Rotor cross section illustrating momentum balance terms (Courtesy GE Oil and Gas)

Figure 1.6 Force giving rotation

If we have constant angular velocity ω, as we do here, then ut is constant, and the time derivative is of the mass, giving us the mass flow rate:

τ = r ut (ρQ (1.9)

However, ut is the tangential velocity of the fluid, not of the impeller. The next task is therefore to determine the velocities of the fluid. The velocity diagram in Fig. 1.6 shows how the actual velocity u is made up from the vector combination of the vane velocity uv and the impeller tangential

Force

Radius


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velocity uti . If we use index 2 for the outer edge of the impeller, the diagram shows the velocity for the fluid leaving the impeller.

Figure 1.7. Velocity diagram at the impeller rim

From the diagram we get the fluid tangential velocity utf from the triangles

uft = uti – uv cos β = ωr – uv cos β (1.10)

since the tangential velocity of the impeller is ω r. Likewise, the radial velocity ur of the fluid as it leaves is

ur = uv sin β

The flow rate Q of the fluid is the outer area times the radial flow. If we call the outer impeller width ( flow ”thickness”) for b we get

Q = 2πr b ur = 2πr b uv sin β (1.11)

The power of the impeller is by definition torque * angular velocity. However, the power used on the liquid is equal to the difference between torque * angular velocity leaving minus torque times angular velocity entering. This is analogous to a difference in kinetic energy out minus kinetic energy in. Thus, the theoretical pump power is , from eq, (1.9)

P = ρQ ω (ro utf – ri utfi) (1.12)

To find an expression for the head, we get from eq. (1.4)

h = ω( ro utfo – ri utfi)/ g (1.13)

Eqs (1.13) and (1.14) are named the Euler turbomachinery equations.

Usually, the velocity of the fluid entering the impeller is much lower than for the fluid leaving it, and it is a very good approximation to neglect the term for the entering fluid in eq. (1.12) and also then of course in eq. (1.11) for the power.

Thus, combining eqs. (1.12) and (1.13) and also neglecting the inlet term we get for the head

∆ h = g

ωro ( ωro - uv cos β) (1.14)

Solving for uv from eq. (1.14) then gives

h = βπ

ωωtgbg

Q

g

r

2

222 − (1.15)

If we want to introduce n, the number of revolutions per second ( not the usual rpm = rev. per minute), we have


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ω = 2πr2 n/r2 = 2 π n (1.16)

giving

h = β

πtgbg

Qn

g

nD −222

(1.17)

where D is the impeller diameter. Since from eq. (1.4) P = ρQgh, we also get an expression for P from eqs. (1.15) and 1.11).

Eq. (1.16) tells us that the head h does not depend on the density of the liquid. One consequence of this is that if a centrifugal pump is started with air in the pipes and the pump, the head of air will be the same as for the liquid, but the corresponding pressure will be so small that the pump will not suck liquid. Thus, the pump must be filled with liquid before starting, it must be primed. There exists so-called self-priming centrifugal pumps, where some liquid is stored in the pump and used for the priming.

1.4. Affinity relations

The second term in eq. (1.17) is smaller than the first, and to derive the similarity relations it is common to neglect that term, equivalent to setting β= 90 o.

For geometrically similar pumps we have that b is proportional to D. If we assume correctly that the impeller is filled with liquid, the flow rate Q will be proportional to the liquid in the impeller which is proportional then to D3 times the rotational speed or to n. The power is proportional to density, flow rate and head. We then get from eq. (1.17) that

(1.18)

These relations can be used over small ranges to predict changes in pump behavior, say from an increase in revolutions or a decrease in D.

1.5. Efficiency

The power given above is again the power without friction or losses. The flow losses come from these sources


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• Impeller recirculation of liquid • Friction losses from the liquid flow • ”Shock” losses from a mismatch between the blade angles and the inlet flow direction

The first loss is most pronounced at low flow rates, the second rises monotonically with the flow rate, and the third is most important at high flow rates. The resulting overall efficiency will thus be low at low flow rates and at high flow rates, and have a highest value in between somewhere. In general, large pumps have higher efficiencies than small pumps.

In addition there are mechanical losses in the pump itself from friction in the bearings etc.

Finally, the efficiency of the electric motor must be taken into account separately in deciding how large a motor is needed, as this efficiency is not given by the pump manufacturers.

1.6. Principle

The equations above for head and power without taking losses into account will give connections between the variables that are not correct over the full range.

One example is the effect of the blade angle β. According to eq (1.17), the head should increase with Q for β>90 o, since tg β <0 then. For β=90 o head is independent of Q since the second term in (1,.17) becomes 0, and for β<90 o the head should fall with increasing Q. For real pumps, this last phenomenon occurs at low flow rates, but then the head falls with increasing Q for all values of β. A head curve that first rises and then falls may give unstable operations, and are most often avoided, and thus the blades are most often curved backwards (β<90o). From eq. (1.17) the head should vary linearly with Q. For real pumps the pump characteristic (h vs Q) is below the straight line.

To determine the real head and power, tests are run for each pump, usually with water. The result is then given as diagrams by the manufacturers. In principle, the curves for head, power and efficiency are as shown in the figure below:


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Figure 1.8. Performance curves for a centrifugal esp (electrical submersible pump). (Courtesy GE Oil and Gas)

The units on the axes must of course be different for power, head and efficiency. The diagram is given for a fixed rotational speed.

The curve for the efficiency η follows the pattern described above. Ideally, the pump should be operated at the point of maximum efficiency, called Best Efficiency Point (BEP). In a process with varying conditions this may not be even theoretically possible, but certainly the pump should be in the vicinity of BEP most of the time, or a different pump should be chosen. Using manufacturers guidelines should ensure a proper match between pump and operating requirements, at least if the latter are properly determined in advance.

1.7. Pump selection diagrams

For each pump model the manufacturer usually will deliver a range of impellers (3-5 typically) with different diameters. For each pump then, a combination of biggest and smallest impeller sizes with suitable operating range will give an area in the h-Q diagram as shown below


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Figure 1.9. Recommended range for a particular pump.

The manufacturers will produce many different sized (different casing etc) pumps within each series. To simplify the selection, the range of all the relevant pumps are shown as areas of the kind given in Fig 1.7, but in one diagram. Note that the maximum efficiency is typically at the “knee” in the curves and falls as the operating point deviates from that value.

1.8. Stage Count

The number of individual stages can be adjusted to yield additional head. It may not always be possible to meet our requirements with one single-stage pump. For larger flow rates, we can operate pumps in parallel. For higher head, we can in principle operate pumps in series. However, there are many standard multistage pumps, i.e. more than one impeller on the same shaft, giving higher head with one pump. This also true for compressors, but there are some physical limitations which affect compressors.

Figure 1.10 Multistage rotors mounted on drive shaft.

The image in Fig 1.11 describes this variation for an ESP application. (Courtesy GE Oil and Gas)


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Figure 1.11 Multistage performance curve for an ESP pump (Courtesy GE Oil and Gas)

1.9. Performance diagrams

For each pump designation ( one particular choice in the selection diagram ) the manufacturer will further provide a detailed performance diagram, like the one shown. Curves for equal efficiences are also shown. The lower diagram gives the power needed for each impeller, flow and efficiency combination from the upper diagram. Sometimes power will be given by iso-power lines in the upper diagram instead.


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Figure 1.12 Centrifugal Pump Performance Curve

1.10. Pump speed

You will note that the pump selection diagram gives the same types of pumps for two different, but fixed, rotational speeds, while the detailed performance diagram is for just one particular constant speed. The reason for this is that most pumps are driven by asynchronous three-phase electric motors, which will have a more or less constant speed depending on their number of poles.

Very large pumps may be driven by steam or gas turbines or by variable speed electric motors. If high speed is needed it is of course possible to employ a gear.

1.1 Effect of viscosity The pump diagrams so far are based on low-viscosity liquids, and come from tests with water. However, increasing the viscosity will decrease the perfomance of a centrifugal pump.


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Figure 1.13 Centrifugal Pump Performance with comparison between process fluid and seawater.

For viscosity ratios above about 300 a positive displacement pump will normally be chosen instead. The manufacturers will provide more detailed guidance on how to adjust for viscosity.

If we go back to eqs. (1.16) for the head h and (1.18) for the similarity relations, and neglect the influence of the angle β ( equal to setting β = 90o), we can write the following proportionalities keeping the correct units (dimensions)

gh ∝ n2 D2 and Q ∝ n D3

2/32/34/3)( Dngh ∝ 2/32/32/1 DnQn ∝

Thus, the following expression is dimensionless

4/3

2/1

)(gh

QnN s = (1.19)


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and Ns is called the specific speed.

If we plot the pump characteristic (h vs. Q) as one curve in a h-Q diagram, and plot the downstream pressure drop (expressed as pressure head or just head), we get two curves looking in principle like those in figure 1.11 below: Which is well understood by both subsea engineers and turbo-machinery staff. The system curve represents the energy lost due to flow and elevation changes. The intersection between the two curves is the operating point

Figure 1.16 Pump characteristic curve and system curve.

If we increase the pressure drop, say by closing a valve a bit, Q decreases and h increases as the system curve shifts upwards.

Figure 1.17 Change in operating point with change in arrival pressure.

The practical consequences are: Not only must we choose a pump that delivers the pressure and flow rate we need, but we must also check in advance that the system curve (the downstream


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pressure drop) is such that the pump operates where we would like it to operate, and let that influence our choice of pump.

The pressure in the liquid goes down a bit at the entrance to the impeller, and rises some towards the rim. Under certain conditions, vapor may form at the lowest pressure, and the vapor bubbles will collapse again after a short passage. This collapse is very abrupt, and with many bubbles we get various operating problems, which in increasing order of seriousness are

• noise • loss of efficiency • vibration • small bits of the metal is torn away (pitting).

The bubble formation and collapse is called cavitation, and it is clearly something we want to avoid.

To avoid cavitation pump manufacturers will specify a minimum pressure at the pump inlet. The specification is given as head, and has been given a strange name which may confuse: Net positive suction head, NPSH (netto positiv sugehøyde).

It is then up to the user of the pump to make sure the actual pressure is at least as great as the required NPSH. The available pressure as head is then sometimes called available NPSH (NPSHa) and the requirement is called required NPSH (NPSHr) To check the available NPSH one has to make an ordinary pressure drop calculation along the suction line, and then subtract the vapor pressure at the temperature of the inlet conditions. Some sources recommend an addition to the required NPSH to have a safety margin, and it may be smart to check with the manufacturer about this too before ordering.

The middle diagram in the pump performance page is NPSH for that particular pump, given only for the largest and the smallest impeller.

Some pumps made for delivering very high pressure (head) may have such high requirement for NPSH that a so-called booster pump is needed, to increase the pressure in the feed to the second pump.

Like other equipment that is used in large numbers, such as tubular heat exchangers, centrifugal pumps are generally manufactured according to some standard. Using a standard means that many details of the pump and not least how it is connected to flanges are the same no matter who the manufacturer is. There is no point in going into details here, it is enough for now to realize that the standards exist and should be used.


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2. Dry-gas Compressors

There are two main types of dynamic compressors which are used in the subsea industry (with a third hybrid). These are illustrated below along with the hybrid helicoaxial device.

Figure 2.1 Illustration of centrifugal and axial compressors alongside the helicoaxial hybrid.

Compressor performance is treated similarly to that for pumps, but there is one major exception coming from the fact that the fluids are compressible. That being that the adiabatic Bernoulli’s equation becomes

0)()(2

1 22 =+++−+− ∫ wf

P

P

oiio hhdP

hhguuo

iρ

(2.1)

Which is more difficult to handle because of the fact that the integral path is neither isothermal nor simple.


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The methodology can be exemplified using ideal gas behaviour. From introductory thermodynamics, we should remember that we have a relationship which can be used to find the temperature as we compress the gas:

(2.2)

This and the definition for enthalpy allow us to avoid the volume integral and use the first law instead:

(2.3)

(2.4)

Which allows for another simplification given that

(2.5)

or

(2.6)

Which gives a convenient form but is not always relevant because we need a more complicated relationship to describe the PVT behaviour of the fluid. The equations most commonly used are cubic-equations-of-state with relationships like

SRK PR

)( bvv

a

bv

RTP

+−

−=

b) - b(v + b) + v(v

a -

b - v

RT = P

P

RT 0.08664 = bc

c c

c

RTb = 0.0778 P

*

1

212

1

2

1

2* lnlnpC

R

p P

PTT

P

PR

T

TC

=⇒

=

∫∆=2

1

*T

T

dTCHWp

−

=−= 1)(

*

1

21

*12

* pC

R

pp P

PTCTTCW

RTPVwhileRCCandC

Cvp

v

p =+== ***

*

γ

−

−=

−

11

1

1

211

γγ

γγ

P

PVPW


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α P

TR 0.42748 = ac

2c

2

⋅ 2 2

c

c

R T a = 0.45724 P

a

])T - m(1 + [1 = 2rα ])T - m(1 + [1 = 2

rα

m = 0.48508 + 1.55171 ω – 0.15613 ω2 m = 0.3746 + 1.54226 ω - 0.26952 ω2

(2.6)

It is possible to expand these equations for use with compressor sizing and this was indeed the state of the art for many years until it became quite clear that a combined approach gave a method that was consistent and also (for the most part) described a system where the efficiency did not change with in the compressor (the efficiency was roughly independent of stage number). The method looks very much like the ideal gas law, but considers a weaker constraint on the PV product.

If the relationship

constant=nPV (2.7)

Then a relationship like that for the ideal gas law is defined

(2.8)

Where n is derived iteratively by fitting the volumetric behaviour over each stage

2.1 Affinity Terms

These are identical to those for a centrifugal pump

2.2 Efficiency

The major causes for efficiency losses with in a centrifugal compressor are similar to those in a centrifugal pump

• Heat losses

• Frictional losses due to high velocity flow in the diffusers

( )

−

−=−=

−

11

11

1

21112

n

n

wp P

P

m

RTZ

n

nHHW

η


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• Leakages

• Mechanical losses

• PVT model inaccuracies

GE reports that the diffuser frictional losses become significant when the compression ratio is large or the number of stages is high. This because the velocities in the final stages become quite high. (The effect is that the polytropic efficiency is not constant). The figure below describes why the equation-of-state model is important. In essence, the discrepancy between the behaviour of the various models results in a difference in polytropic efficiencies. The figure also illustrates why it is important to consider the use of stage-wise integration. The value for n may be constant for individual stages, but in this example, not over the whole pressure range.

Figure 2.2 Comparison of Z-factors


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3. Multiphase Pumping Multiphase boosting poses other design challenges. For liquid-liquid systems, the liquid boosting model is somewhat correct, but the fact that viscosity effects the pump performance means that the effect viscosity of emulsion forming systems needs to be considered. This is very important for systems which are already viscous. The result is that there exists multiple separate solution sets for liquid rich systems.

Figure 3.1 Illustration of multiphase boosting technologies.

Low Viscosity Liquid High Viscosity Liquid

Low GVF (<5%) Centrifugal/Helicoaxial Centrifugal/TwinScrew

Moderate GVF (5-50%) Helicoaxial Helicoaxial

High GVF (50-95%) Helicoaxial/Twinscrew TwinScrew

Head requirement also comes into play and currently the twinscrew and centrifugal pumps have the lead on pressure generation capability. The issue which helicoaxial pumps have with high viscosity mixtures is that like axial compressors, they generate swirl downstream which inturn leads to choke-flow in the discharge.

Because of the presence of all three phases, the design models are more complex and revert to classic second-law analysis. The system cannot be treated with a polytropic efficiency as the heat capacity of the liquid is high enough to result in nearly isothermal flow which yields an ill-conditioned equation set. The analysis is instead based on the classic Entropy-Enthalpy analysis illustrated below for pure steam.


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Figure 3.2 Classical method for calculating duty and discharge conditions

The issue here is that it is now related to the PVT behaviour of a more complex fluid which has a more complex characterization. The fluids at the inlet of the pump are similar to those in a reservoir with a gas cap. In this case, they co-exist at the intersection of the two curves but with the pump path taking the two fluids upwards (and probably to the right for increasing temperatures). The question is whether the fluids will convert to single phase within the pump (as with nitrogen-lubrication oil experiments) or whether it will remain two-phase for some distance downstream of the contact point (as is a concern in gas-injection projects)


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Figure 3.3 Associated Phase PT envelopes

At any rate, the calculation methodology becomes similar to that for a turbo-expander and is available in most process simulators.

3.1 Efficiency

The major causes for efficiency losses with in a multiphase boosters compressor are similar to those in a centrifugal pump

• Heat losses (at high GVF’s where cooling is poor)

• Slip (in helico-axial and twin-screw designs)

• Mechanical losses: in particular gear windage in twin-screws

• PVT model inaccuracies

The mechanisms for slip and internal leakage are similar to those for compressors and pumps, but because the units have to be designed for both high and low viscosities the performance curves of multiphase phase devices can vary.


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Figure 3.4 Flow and leakage paths in a twinscrew pump.

According to the models published at Turbolab.tamu.edu, this form of slip is treated as laminar flow through a micron gap and will break down at high rates. It should be noted that the models for multiphase flow in twinscrews assume that the liquid phase migrates to the outside because of cyclonic effects in the rotation. This may not be the case as the main flow direction is normal to the direction of rotation. Xu treats the slip velocity as being related to the differential pressure only

57.0)(* stageslip Pku ∆= (3.1)

Where the stage is the Peak-to-Peak volume. The volumetric flow capacity is otherwise directly proportional to the stage volume and the frequency.

Figure 3.5 Effect of slip on generated head in a twin screw pump


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Figure 3.6 Slip and Recirculation paths in an axial compressor / helicoaxial pump

Figure 3.7 Comparison of capacity curves for axial and centrifugal compressors.

3.2 Surge and Stall

All of the dynamic compressor types are prone to flow instabilities at both high and low head operations. As the operating point moves to the left in Figure 2.6, the system becomes very sensitive to changes in differential head. This is noted in a form of vibration known as surge.


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Axial compressors are prone to blade stall as the blades fail to generate head. Modern control systems are designed to protect the compressor from both conditions.

Figure 3.8 Surge and Stall Control Lines

Figure 3.9 Typical Compressor Protection Scheme using antisurge recycles.


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3.3 Cavitation and NPSH

Cavitation is a phenomenon associated with boosting and level control and is associated with the flow rate from the liquid source to the pump and the pressure drop in the piping. Most literature references describe the limit in terms of the cavitation number:

2*

)(*2

u

PPK v

ρ−

= (3.2)

This formula should be familiar to those working in erosion (and is in fact the source of the format for API RP 14E) which is aimed directly at the process illustrated below: designing a piping solution for a liquid flowing by differential pressure controlled flow and little elevation change. The classic topsides solution is to use available height to increase the pressure at the pump inlet by putting the pumps on the so-called cellar deck.

Figure 3.10 Classic level control solution using an export pump.

Figure 3.11 Illustration of effect of loss of NPSH on pump performance

3.4 Erosion and Deposits


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Droplets and sand particles in the fluids can cause erosion damage to the booster. This is an issue for all types of devices but some are more prone to damage than others.

• Twin-screw pumps are prone to wear between the casing and the rotor. This increases slip and can perforate the casing.

• Centrifugal pumps/compressors are prone to erosion in the diffusers. This increases circumferential slip, threatens the casing and can cause imbalance.

• (Helico)Axial pumps/compressors are prone to blade wear. This causes slip, imbalance and a can lead to blade failures.

Deposits are an issue for all three and are known to cause notable efficiency losses in axial and centrifugal devices. (Twin Screw devices are somewhat self cleaning). Because deposits are known to change the stall speed this should also provide a complication on axial compressor control. The surge speed is expected to change for both dynamic devices. The problem is illustrated below showing asphaltene deposits on a downhole flow measurement tool.

Figure 3.12 Asphaltene deposits on a PLT spinner.


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4. Subsea Operational Issues 4.1 Start-up and shut down

A topsides process is designed to provide a high quality product in a safe and robust manner. This means that the process looks like that illustrated below: These facilities include devices intended to protect the rotating equipment.

Threat Device Mitigation

Sand Pumps/valves Separators and strainers

Droplets (Carry-over) Compressors Scrubbers and internals

Slugging Pumps/compressors Inlet Separator Volume

Deposits All Chemical Treatment

Figure 4.1 Atypical topsides process

This complexity means that a compressor or pump is conserved with its native fluid during a shut-down or start-up. This is not necessarily the case for a subsea application the fluids will segregate in the vertical piping so that the equipment may have to tolerate very different conditions than those for which they are designed. (Ref API 676)

Device Normal Condition Transient Condition


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Pump Liquid filled Gas-filled

Gas slugs

Sand-free Sand «slugs»

Discharge flow Blocked discharge

Compressor Gas-filled Liquid-filled

Liquid slugs

Carry-over

Sand-free Fines

Electric Driver Clean dielectric Gas-contaminated

Sour

Water contaminated

Fines

Bearings Clean Lubricant or Gas Sand

Liquids

In-fact API 676 includes an optional requirement for ( potentially destructive). In fact one issue (driver protection) is not even an issue on the topsides system. This means that a subsea booster system ends up as being quite complex (despite operator wishes to the contrary)

Figure 4.2 Block diagram of a subsea system.

This is in fact illustrated by the subsea portion of the Ormen Lange test facility. The compressor station will be tested for tolerance to the conditions described above. The reader will also not the complexity of the electrical solution caused by the long step-out. A similarly complicated facility is installed outside the pit to provide facilities to emulate slugging, sand production, etc.


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Figure 4.3 Ormen Lange Test Pit, Items in the green block are new.


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5. References and Relevant Reading Friedemann J.D and Løvland, J.. «Offshore Processing», NTNU, Trondheim Norway, Jan 2006.

Brennen, C. E. «Hydrodynamics of Pumps», Oxford University Press, London, UK1994

Hanlon, P.C. ed. Compressor Handbook, McGraw-Hill, NYC, NY 2001

Xu, J. Modeling of Wet Gas Compression in Twin-Screw Multiphase Pump, PhD Thesis, Texas A&M University, College Station TX 2008.


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6. Selection of Relevant Standards ANSI B73.1 Specification for horizontal end suction centrifugal pumps for chemical process

ANSI B73.2 Specification for vertical in-line centrifugal pumps for chemical process

API 610 / ISO13709 Centrifugal pumps for petroleum petrochemical and natural gas industries.

API 617/ISO 10439 Axial and centrifugal compressors and expander-compressors for petroleum, chemical and gas industry services.

API 618/ISO 13707 Reciprocating compressors for petroleum, chemical and gas industry services.

API 685/ISO 13709 Sealless Centrifugal pumps for Petroleum, petrochemical and gas industry process service.

API 676 Positive Displacement Pumps – Rotary

ISO 10440 Petroleum, petrochemical and natural gas industries – Rotary-type positive displacement compressors – Part 1: Process compressors

1. Centrifugal Pumps - Forsiden - Universitetet i Oslo · 1. Centrifugal Pumps ... Since pumps and...

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