& @¢¢ ¢ ¢@¢¢
`
`$6¢¢¢€8¢6€
`>68888
`
`$
`
`IV-728
`
`Petroleum Engineering Handbook—Vol. IV
`
`Jet Pump Nomenclature
`
`Nozzle
`
`Throat
`
`Diffuser
`
`Red—high pressure
`Blue—production
`Purple—mixed powerfluid and production
`Fig. 14.11—Howthe jet pump works.
`
`low pump intake pressures, and this must be considered in design calculations.
`the throat at
`Also, because of the nature of their performance curves, the calculations used for installation
`design are complex anditerative in nature and are best handled by computers. Their overall
`energy efficiencies are low, which may lead to high energy costs; despite these limitations,
`their reliability,
`low maintenance costs, and volume capability make them attractive in many
`wells, and their use has increased since commercial introduction in the early 1970s.
`
`larger-diameter nozzles and throats would
`Intuitively,
`14.3.3. Performance Characteristics.
`seem to have higher flow capacities, and this is normally the case. The ratio of the nozzle area
`to the throat area is an important variable because it determines the trade-off between produced
`head and flow rate. Fig. 14.11 shows a schematic of the working section of a jet pump. If, for
`a given nozzle, a throat is selected such that the area of the nozzle, A,,
`is 60% of the area of
`the throat, 4,, a relatively high-head, low-flow pump will result. There is a comparatively small
`area, A,, around the jet for well fluids to enter. This leads to low production rates compared to
`the power-fluid rate, and because the energy of the nozzle is transferred to a small amount of
`production, high heads develop. Such a pumpis suited for deep wells with high lifts, and sub-
`stantial production rates can be achieved if the pump is physically large, but the production
`rate will always be less than the power-fluid rate.
`If a throat is selected such that the area of the nozzle is only 20% of the area of the throat,
`much more flow area around the jet is available for the production. However, because the noz-
`zle energy is transferred to a large amount of production compared to the power-fluid rate,
`lower heads will be developed. Shallow wells with low lifts are candidates for such a pump.
`Any number ofsuch area combinations is possible to match different flow and lift require-
`ments. Attempting to produce small amounts of production compared to the power-fluid rate
`
`GDI Ex. 1008
`Page 63 of 91
`
`

`

`Chapter 14-Hydraulic Pumping in Oil Wells
`
`IV-729
`
`with nozzle/throat-area ratio of 20% will be inefficient because of high-turbulent mixing losses
`between the high-velocity jet and the slow-moving production. Conversely, attempting to pro(cid:173)
`duce high production rates compared to the power-fluid rate with a nozzle/throat-area ratio of
`60% will be inefficient because of high friction losses as the produced fluid moves rapidly
`through the relatively small throat. Optimal ratio selection involves a trade-off between these
`mixing and friction losses.
`As a type of dynamic pump, the jet pump hns characteristic perfonnance curves similar to
`those of an ESP. A family of perfonnance curves is possible, depending on the nozzle pressure
`supplied to the pump from the surface. Different sizes of throats used in conjunction with a
`given nozzle size give different perfonnance curves. The curves are generally fairly flat, espe(cid:173)
`cially with the larger throats, which makes the jet pump sensitive to changes in intake or
`discharge pressure. Because variable fluid mixture densities, gas/liquid ratios, and viscosity af(cid:173)
`fect the pressures encountered by the pump, the calculations to simulate performance are
`complex and iterative in nature and lend themselves to a computer solution.
`
`14.3.4 Cavitation in Jet Pumps. Because the production must accelerate to a fairly high veloc(cid:173)
`ity (200 to 300 ft/sec) to enter the throat, cavitation is a potential problem. The throat and
`nozzle flow areas define an annular flow passage at the entrance of the throat. The smaller this
`area is, the higher the velocity is of a given amount of produced fluid passing through it. The
`static pressure of the fluid drops as the square of the velocity increases, declining to the vapor
`pressure of the fluid at high velocities. This low pressure causes vapor cavities to fonn, a pro(cid:173)
`cess called cavitation. This results in choked flow into the throat, and production increases are
`not possible at that pump-intake pressure, even if the power-fluid rate and pressure are in(cid:173)
`creased. Subsequent collapse of the vapor cavities, as pressure is built up in the pump, may
`cause erosion known as cavitation damage. Thus, for a given production flow rate and pump
`intake pressure, there is a minimum annular flow area required to keep the velocity low
`enough to avoid cavitation. This phenomenon has been the subject of numerous investigations~
`the most notable being that of Cunningham and Brown,21 who used actual oilwell pump de(cid:173)
`signs at the high pressures used in deep wells.
`The description of the cavitation phenomenon suggests that if the production flow rate ap(cid:173)
`proaches zero, the potential for cavitation will disappear because the fluid velocities are very
`low. Under these conditions, however, the velocity difference between the power-fluid jet and
`the slow-moving production is at a maximum, which creates an intense shear zone on the bound(cid:173)
`ary between them, generating vortices, the cores of which are at a reduced pressure. Vapor
`cavities may form in the vortex cores, leading to erosion of the throat walls as the bubbles
`collapse because of vortex decay and pressure rises in the pump. Although no theoretical treat(cid:173)
`ments of this phenomenon have been published, it has been the subject of experimental work,
`which has led to the inclusion, by suppliers, of potential damage zones on their published per(cid:173)
`formance prediction plots. This experimental correlation predicts cavitation damage at low flow
`rates and low pump-intake pressures before the choked flow condition occurs. Field experience
`has shown, however, that in most real oil wells, the erosion rate in this operating region is
`very low, probably because of produced gas cushioning the system by reducing the propagation
`velocity of the bubble-collapse shock waves. It is generally agreed that this phenomenon is of
`concern only in very-high-water-cut wells with virtually no gas present. Under these condi(cid:173)
`tions, cavitation erosion has been observed even at very low production rates; however, if a jet
`pump is operated near its best efficiency point, the shear vortices are a distinctly second-order
`effect in the cavitation process.
`
`(cid:141)
`
`14.3.5 Nozzle and Throat Sizes. Each manufacturer has different sizes and combinations of
`nozzles and throats. Manufacturers A and B increase the areas of nozzles and throats in a geo(cid:173)
`metric progression (i.e., the flow area of any nozzle or throat is a constant multiple of the area
`
`GDI Ex. 1008
`Page 64 of 91
`
`

`

`¢¢e¢
`
`é&@¢é@¢é@e¢¢¢
`
`@ é
`
`
`TABLE 14.4—NOZZLE AND THROAT SIZES
`
`
`
`Nozzle
`Area
`Throat No.
`Area
`
`Petroleum Engineering Handbook—vVol. IV
`
` IV-730
`
`
`0.0024
`0.0031
`0.0039
`0.0050
`0.0064
`0.0081
`0.0103
`0.0131
`0.0167
`0.0212
`0.0774
`0.0346
`0.0441
`0.0562
`0.0715
`0.0910
`0.1159
`0.1476
`0.1879
`0.2392
`
`1
`2
`3
`4
`5
`6
`7
`8
`9
`10
`11
`12
`13
`14
`15
`16
`17
`18
`19
`20
`
`0.0064
`0.0081
`0.0104
`0.0131
`0.0167
`0.0212
`0.0271
`0.0346
`0.0441
`0.0562
`0.0715
`0.0910
`0.1159
`0.1476
`0.1879
`0.2392
`0.03046
`0.3878
`0.4938
`0.6287
`
`Manufacturer A:
`1
`2
`3
`4
`5
`6
`7
`8
`9
`10
`11
`12
`13
`14
`15
`16
`17
`18
`19
`20
`Manufacturer B:
`0.0060
`4
`0.0024
`1
`0.0077
`2
`0.0031
`2
`0.0100
`3
`0.0040
`3
`0.0125
`4
`0.0052
`4
`0.0167
`5
`0.0067
`5
`0.0215
`6
`0.0086
`6
`0.0278
`7
`0.0111
`7
`0.0359
`8
`0.0144
`8
`0.0464
`9
`0.0186
`9
`0.0599
`10
`0.0240
`10
`0.0744.
`11
`0.0310
`11
`0.1000
`42
`0.0400
`12
`0.1292
`13
`0.0517
`13
`0.1669
`14
`0.0668
`14
`0.2154
`15
`0.0863
`15
`0.2783
`16
`0.1114
`16
`0.3594
`17
`0.1439
`17
`0.4642
`18
`0.1858
`18
`0.5995
`19
`0.2400
`19
`0.7743
`20
`0.3100
`20
`
`
`of the next smaller size). Manufacturer B’s factor is 1.29155, and Manufacturer A’s factor is
`4/m = 1.27324. The system of sizes offered by Manufacturer C uses a similar geometric progres-
`sion concept but does not use the same factor over the total range. In the smaller sizes, where
`the change in horsepowerper size is small,
`the rate of increase in area is more rapid than in
`the systems of Manufacturers A and B. In the larger, higher-horsepowersizes, the percent in-
`crease in size is less rapid than in the other systems to limit
`the incremental
`increase in
`horsepower. The sizes offered by Manufacturer C cover a slightly larger range than those of
`Manufacturers A and B. The sizes from these manufacturers are listed in Table 14.4. The max-
`imum sizes of nozzles and throats that are practical in pumps for a given tubing size depend
`on the fluid passages of the particular pump, BHA, swab nose, and standing valve. Single-seal
`pumps cannot use nozzles as large as those practical
`in higher-flow, multiple-seal pumps. In
`general, nozzles larger than 0.035 in.? in flow area are used only in pumps for 2%4- and 3%-in.
`tubing.
`
`GDI Ex. 1008
`Page 65 of 91
`
`

`

`Chapter 14—Hydraulic Pumping in Oil Wells
`
`IV-731
`
`
`
`TABLE 14.4—NOZZLE AND THROAT SIZES (Continued)
` Nozzle
`Throat No.
`
`Manufacturer C:
`DD
`cc
`BB
`
`0.0016
`0.0028
`0.0038
`
`000
`00
`0
`
`Area
`
`0.0044
`0.0071
`0.0104
`
`V2EBrFARO7-TOMmMOOW>Y
`
`
`
`if
`
`Page 66 of 91
`
`
`
`The strict progression used by Manufacturers A and B establishes fixed area ratios between
`the nozzles and different throats. A given nozzle matched with the same numberthroat always
`gives the same area ratio: 0.380 in Manufactures A’s system and 0.400 in Manufacturer B’s
`system (Table 14.4). This is called the A ratio. Successively larger throats matched with a giv-
`en nozzle give the B, C, D, and E ratios. In the systems of Manufacturers A and B, the nozzle
`size and ratio designate the size of a pump. Examples are 11-B, which is a No. 11 nozzle and
`a No. 12 throat, and 6-A, which is a No. 6 nozzle and a No. 6 throat.
`Because the size progression for the nozzles and throats in Manufacturer C’s systemis not
`constant over the whole range, the nozzle/throat combinations do not yield fixed ratios. Howev-
`er, the ratios that result cover the same basic range as the other two systems. The actual ratios
`are listed in Table 14.5. In Manufacturer C’s system, the nozzle and mixing tube (throat) sizes
`designate the size of a pump. An example is C-5, which are the size C nozzle and the No. 5
`throat. This combination has an area ratio of 0.32. The annular flow areas of Manufacturer C’s
`jet pumpsused in cavitation calculations are also included in Table 14.6. The annular areas for
`Manufacturers A and B’s jet pumpsare listed in Tables 14.6 and 14.7.
`The most commonly used area ratios fall between 0.235 and 0.400. Area ratios greater than
`0.400 are sometimes used in very deep wells with high lifts or when only very lowsurface
`operating pressures are available and a high head regain is necessary. Area ratios less than
`0.235 are used in shallow wells or when very low BHPs require a large annular flow passage
`to avoid cavitation. The smaller area ratios develop less head but may produce more fluid than
`is used for powerfluid (F,,> 1.0). Where the curves for different area ratios cross, the ratios
`have equal production and efficiency; however, different annular flow areas (4,) may give
`them different cavitation characteristics.
`
`14.3.6 Jet-Pump Application Sizing. The widespread current use of jet pumps can be credit-
`ed to the advent of computer programs capable of making the iterative calculations necessary
`
`
`
`GDI Ex. 1008
`
`

`

`
`
`
`\V-732
`
`-t+——
`
`Nozzle
`
`DD
`
`CC
`
`BB
`
`A
`
`B
`
`Cc
`
`D
`
`E
`
`F
`
`G
`
`H
`
`I
`
`J
`
`K
`
`L
`
`M
`
`N
`
`P
`
`Throats
`R
`As
`Throats
`R
`As
`Throats
`R
`As
`Throats
`R
`As
`Throats
`R
`As
`Throats
`R
`As
`Throats
`R
`As
`Throats
`R
`As
`Throats
`R
`As
`Throats
`R
`As
`Throats
`R
`As
`Throats
`R
`As
`Throats
`R
`As
`Throats
`R
`As
`Throat
`R
`As
`Throats
`R
`As
`Throat
`R
`As
`Throat
`R
`As
`
`000
`0.36
`0.0028
`000
`0.64
`0.0016
`00
`0.37
`0.0032
`0
`0.53
`0.0048
`0
`0,92
`0.0009
`1
`0.86
`0.0020
`3
`0.74
`0.0054
`4
`0.77
`0.0074
`8
`0.69
`0.0138
`8
`0,68
`0.0206
`10
`0.69
`0.0302
`11
`0.72
`0.0339
`13
`0.71
`0.0515
`15
`0.61
`0.1015
`16
`0.63
`0.1164
`17
`0.66
`0.1164
`18
`0.69
`0.1395
`19
`0.71
`0.1575
`
`00
`0.22
`0.0056
`00
`0.40
`0.0043
`0
`0.37
`0.0065
`1
`0.39
`0.0088
`1
`0.66
`0.0048
`2
`0.65
`0.0086
`4
`0.56
`0.0137
`5
`0.63
`0.0140
`7
`0.59
`0.0217
`9
`0.56
`0.0352
`1
`0.55
`0.0643
`12
`0.59
`0.0597
`14
`0.56
`0.0908
`16
`0.51
`0.1537
`17
`0.52
`0.1787
`18
`0.55
`0.2050
`19
`0.57
`0.2305
`20
`0.59
`0.2670
`
`1
`0.20
`0.0115
`2
`0.20
`0.0150
`3
`0.23
`0.0185
`3
`0.40
`0.0145
`4
`0.39
`0.0191
`6
`0.39
`0.0276
`7
`0.45
`0.0290
`9
`0.39
`0.0490
`11
`0.38
`0.0742
`13
`0.37
`0.1112
`14
`0.40
`0.1309
`16
`0.40
`0.1671
`18
`0.35
`0.2922
`19
`0.36
`0.3460
`20
`0.38
`0.4055
`
`0
`0.27
`0.0076
`1
`0.27
`0.0105
`2
`0.29
`0.0133
`2
`0.50
`0.0094
`3
`0.51
`0.0116
`5
`0.46
`0.0203
`6
`0.53
`0.0212
`8
`0.46
`0.346
`10
`0.47
`0.0510
`12
`0.45
`0.0792
`43
`0.46
`0.0917
`45
`0.48
`0.1349
`17
`0.42
`0.2150
`18
`0.44
`0.2549
`19
`0.45
`0.2961
`20
`0.48
`0.3401
`
`TABLE 14.5—MANUFACTURER C RATIOS AND THROAT ANNULUS AREAS, IN.”
`
`
`~
`
`Petroleum Engineering Handbook—Vol. IV
`
`
`6
`0.21
`0.0357
`7
`0.23
`0.0406
`9
`0.22
`0.0628
`10
`0.25
`0.0722
`42
`0.22
`0.1138
`14
`0.21
`0.1712
`16
`0.21
`0.2467
`17
`0.23
`0.2895
`19
`0.23
`0.4167
`
`5
`0.25
`0.0285
`6
`0.27
`0.0330
`8
`0.27
`0.0464
`9
`0.30
`0.0564
`11
`0.26
`0.0860
`13
`0.26
`0.1320
`15
`0.25
`0.1945
`16
`0.27
`0.2272
`18
`0.28
`0.3256
`20
`0.24
`0.4928
`
`4
`0.30
`0.0219
`5
`0.32
`0.0257
`7
`0.33
`0.0354
`8
`0.36
`0.0420
`10
`0.33
`0,848
`12
`0.31
`0.1000
`14
`0.30
`0.1504
`45
`0.33
`0.1750
`17
`0.34
`0.2493
`19
`0.29
`0.3833
`20
`0.30
`0.4556
`
`11
`0.20
`0.0954
`
`i
`
`&tf
`
`l
`fe
`
`ێ>
`
`=
`
`ee
`
`ee
`
`for application design. Jet-pump performance depends largely on the pump discharge pressure,
`which in turn is strongly influenced bythe gas/liquid ratio, F,,, in the return column to the
`surface, higher values of F,; lead to reduced pump discharge pressure. Because the jet pump is
`
`GDI Ex. 1008
`Page67 of 91
`
`

`

`Chapter 14—Hydraulic Pumpingin Oil Wells
`
`IV-733
`
`Nozzle
`
`
`
`0.0144
`0.0163
`0.0233
`0.0296
`0.0377
`0.0481
`0.0812
`0.0779
`0.0992
`0.1264
`0.1608
`0.2046
`0.2605
`0.3316
`0.4223
`0.5377
`
`
`
`
`
`
`
`>
`
`Se
`
`(a
`
`=
`
`-
`&
`
`am
`
`«>
`.
`
`>
`
`(~
`
`@
`
`se
`
`@
`
`“>
`
`sm
`“
`
`se
`
`-
`
`S
`.
`
`ES
`
`ip
`€&>
`
`om
`
`&
`
`>
`
`a
`
`“y
`
`0.0057
`0.0040
`4
`0.0073
`0.0050
`0.0033
`2
`0.0093
`0.0065
`0.0042
`3
`0.0118
`0.0082
`0.0054
`4
`0.0150
`0.0104
`0.0088
`5
`0.0191
`0.0133
`0,0087
`6
`0.0243
`0.0169
`0.0111
`7
`0.0310
`0.0215
`0.0141
`8
`0.0395
`0.0274
`0.0179
`9
`0.0503
`0.0350
`0.0229
`| 10
`0.0639
`0.0444
`0.0291
`11
`0.0813
`0.0564
`0.0369
`12
`0.1035
`0.0716
`0.0469
`13
`0.1317
`0.09144
`0.0597
`14
`0.1677
`0.1154
`0.0761
`15
`0.2136
`0.1482
`0.0969
`16
`0.2720
`0.1858
`0.1234
`17
`0.3469
`0.2403
`0.1571
`18
`19
`0.2000
`0.3060
`0.4409
`20
`0.2546
`0.3896
`
`
`0.0080
`0.0101
`0.0129
`0.0164
`0,0208
`0.0265
`0.0339
`0.0431
`0.0548
`0.0698
`0.0888
`0.1130
`0.1438
`0.1830
`0.2331
`0.2968
`0.3779
`0.4812
`
`0.0108
`0.0137
`0.0175
`0.0222
`0.0282
`0.0360
`0.0459
`0.0594
`0.0743
`0.0947
`0.1205
`0.1533
`0.1954
`0.2484
`0.3163
`0.4028
`0.5128
`
`
`TABLE 14.7— THROAT ANNULUS AREAS FOR MANUFACTURER B
`
`
`Nozzle
`1
`2
`3
`4
`5
`6
`7
`8
`9
`10
`11
`12
`13
`14
`15
`16
`17
`18
`19
`20
`
`0.0029
`0.0037
`0.0048
`0.0062
`0.0080
`0.0104
`0.0134
`0.0174
`0.0224
`0.0289
`0.0374
`0.0483
`0.0624
`0.0806
`0.1036
`0.1344
`0.1735
`0.2242
`0.2895
`
`0.0036
`0.0046
`0.0060
`0.0077
`0.0100
`0.0129
`0.0167
`0.0216
`0.0278
`0,0360
`0.0484
`0.0509
`0.0774
`0.1001
`0.1287
`0.1666
`0.2155
`0.2784
`0.3595
`0.4643
`
`+
`.
`.
`ratio (GOR) and on the
`inherently an OPF device, F,,; depends on the formation gas/oil
`amount of power-fluid mixed with the production, which in turn depends on the size of the
`nozzle and the operating pressure. As the power-fluid pressure is increased, the lift capability
`of the pump increases, but the additional power-fluid rate decreases F,,, thereby increasing the
`effective lift. Finding a match between the power-fluid rate,
`the pump performance curve and
`the pump discharge pressure, p,
`is an iterative procedure involving successive refined guesses.
`The various suppliers of jet pumps also have developed in-house computer programs for
`application design that are faster than the past calculator routines and incorporate more correla-
`tion for
`fluid properties and the pump discharge pressure. The object of the calculation
`
`GDI Ex. 1008
`Page 68 of 91
`
`€
`

`

`IV-734
`
`Petroleum Engineering Handbook-Vol. IV
`
`WALS
`016C
`
`Design operating parameters
`Target production: i 0,000 B/D
`Power-fluid injection pressure: 4,000 psi
`Power-fluid injection volume 7,859 B/D
`
`Pump intake pressure: 2,200 psi
`Horsepower: 594
`
`-~ 3,000
`t.lJ c..
`~ 2,500
`a..
`::l
`t.lJ
`t.lJ
`~ 2,000
`n.
`~ JS i,500
`
`C
`
`·1,000
`
`500
`
`15,000
`10,000
`Production Rate, 8/D
`
`vv
`
`vv·en lPR
`
`l -2.50U.
`2-- 3,000.
`Fig. 14.12-Performance plot for the jet pump system.
`
`3-
`pf
`4 - 4,000. psig pf
`
`sequence is to superimpose a jet-pump performance curve on the inflow performance relation-
`ship (IPR) curve of the well and to note the intersections that represent the pump performance
`in that particular well. Therefore, a plot of the best estimate of the IPR or productivity index
`(PI) curve of the well is the starting point. An example of a completed performance plot in this
`format is shown in Fig. 14.12.
`
`14.3.7 Calculation Sequence and Supplemental Equations. Fig. 14.13 shows a typical jet-
`pump installation with the appropriate pressures that determine pump operation. Although a
`parallel installation is shown for clarity of nomenclature, the same relationships hold for the
`casing-type installation.
`
`14.4 Downhole Pump Accessories
`
`14.4.1 Swab Cups. A number of accessories are available for downhole pumping systems. Free(cid:173)
`pump systems require swab cups and a standing valve to accomplish the pump-in and pump(cid:173)
`out operations. The swab cups are carried on a mandrel, extending above the pump, which may
`contain a check valve to limit the amount of fluid by passing the pump as it is circulated to
`the surface. If the pump does not enter a lubricator on the wellhead, the check valve may in(cid:173)
`clude a valve bypass that is actuated when the pump enters the wellhead catcher to prevent
`excessive pressure buildup. Two examples of swab cup assemblies are shown in Fig. 14.14. Jet
`pumps usually use the simpler system.
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`GDI Ex. 1008
`Page 69 of 91
`
`

`

` Chapter 14—Hydraulic Pumping in Oil Wells
`
`D =pumpsetting depth,ft
`
`P., = surface operating pressure, psi
`P., = friction in powertubing,psi
`g, = Yradient of powerfluid, psi/ft
`P, =useful powerfluid presure at nozzle, psi
`P, =P..+ Qn — Pi.X psi
`
`P,, = friction in discharge tubing,psi
`g, = gradientofreturnfluid, psi/ft
`P.,, = flow line pressure at wellhead,psi
`Pog = PUMP discharge pressure,psi
`Pag = Gag t Pig t Pay X PS
`
`P = pumpsuction pressure, psi
`
`IV-735
`
`GDI Ex. 1008 Page 70 of 91
`
`Fig. 14.13—Schematic for jet pumping.
`
`14.4.2 Standing Valves. Standing valves are necessary in free-pump systems to create a “U”
`tube and prevent the circulating fluid from flowing back into the reservoir. During pumping
`operations, the standing valve is opened by flow from the formation to the pump suction; when-
`ever the pump is shut down, the standing valve closes. In some cases, the standing-valve ball
`is held open by a small magnet to prevent it from cycling during reciprocating pump-stroking
`reversals. When the downhole pumpis unseated, fluids attempting to flow back into the forma-
`tion washthe ball off the magnet and onto the seat. The standing valve is wireline-retrievable
`and includes a provision for draining the tubing before attempting to pull it. In most cases, the
`standing valve forms the no-go and bottom seal for the pump. Some jet-pump installations,
`however, use high-flow designs that do not serve as a pumpseat. An example of eachtype is
`shown in Fig. 14.15.
`
`14.4.3 Pressure Recorders. To obtain producing BHPs at several different withdrawal rates,
`downhole pressure recorders are often run in conjunction with hydraulic pumps, hung below
`the standing valve. While this arrangement provides not only pressure drawdownbutalso pres-
`sure-buildup data,
`it has the disadvantage of requiring wireline operations to run and retrieve
`the recorder. Some reciprocating pumps can be run with a pressure recorder attached, which
`eliminates the wireline operations but does not permit observation of pressure buildup because
`the recorder is above the standing valve. Virtually all jet pumps can be run with recorders
`attached, and very smooth recordings are obtained.
`
`14.4.4 Dummy Pumps. Dummy pumps are sometimes run to blank off one or more tubing
`strings so that they may be checked for leaks. If the dummy pump has a fluid passage in it,
`the terms “flow-through dummy” or “blanking tool” are often used. These tools are useful for
`acidizing or steaming.
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`IV-736
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`Pump with bypass
`check valve
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`Petroleum Engineering Handbook—Vol. IV
`
`SS
`
`———DIZRANEAT
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`Fig. 14.14—Hydraulic pump swab cup assemblies.
`
`the downhole pump from trash in the well, various
`14.4.5 Screens and Filters. To protect
`types of screens and filters are sometimes run. Because circulating pumps in and out of a well
`may dislodge scale andcorrosion products in the tubing,a starting filter can be attached to the
`swab-cup assembly to filter the power fluid. Because this must be a relatively small filter,
`it
`will eventually plug up, and an automatic bypass arrangement is provided. This system collects
`foreign material during the crucial startup phase with a newly installed pump. For long-term
`
`
`
`GDI Ex. 1008
`Page 71 of 91
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`

`

`Chapter 14-Hydraulic Pumping in Oil Wells
`
`IV-737
`
`Hydraulic
`free-type
`standing
`valve
`
`High-flow-rate
`standing valve
`for large, free
`jet pumps
`
`Fig. 14.15-Standing valves.
`
`operation, power-fluid and pump intake screens or strainers are used, which exclude large-diam(cid:173)
`eter objects that could damage or plug the pump.
`
`14.4.6 Safety Valves. In some areas, subsurface safety valves are required. When a packer is
`set and the BHA is above it, a wireline-retrievable safety valve can be installed between the
`standing valve and the packer to isolate the formation. The safety valve is normally closed
`
`GDI Ex. 1008
`Page 72 of 91
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`

`IV-738
`
`Petroleum Engineering Handbook-Vol. IV
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`unless the pump supplies high-pressure fluid to it by way of a control line run from the main
`power-fluid tubing just above. The pump discharge pressure provides the reference pressure to
`the safety valve. When the pump is on bottom and power-fluid pressure is applied to it, the
`safety valve opens to allow well fluid to enter the pump. Most safety valves will not hold
`pressure from above, so the standing valve is still necessary for circulating the pump in and
`out of the well. Fig. 14.16 illustrates this type of installation.
`
`14.5 Surface Equipment
`
`14.5.1 Surface Pumps. Hydraulic pumping systems have evolved toward the use of relatively
`high pressures and low flow rates to reduce friction losses and to increase the lift capability
`and efficiency of the system. Surface operating pressures are generally between 2,000 and
`4,000 psi, with the higher pressures used in deeper wells, and power-fluid rates may range
`from a few hundred to more than 3,000 BID. While some surface multistage centrifugal pumps
`are rated to this pressure range, they are generally quite inefficient at the modest flow rates
`associated with single-well applications. Multistage centrifugals can be used effectively when
`multiple wells are pumped from a central location. The surface pump for a single well or for
`just a few wells must be a high-head and low-specific-speed pump. Wide experience in the
`overall pumping industry has led to the use of positive-displacement pumps for this type of
`application, and triplex or quintuplex pumps, driven by gas engines or electric motors, power
`the vast majority of hydraulic pump installations. See Fig. 14.17.
`Multiplex pumps consist of a power end and a fluid end. The power end houses a
`crankshaft in a crankcase. The connecting rnc1s are similar to those in internal combustion en(cid:173)
`gines, but connect to crossheads instead of pistons. The fluid end houses individual plungers,
`each with intake and discharge check valves usually spring loaded, and is attached to the pow(cid:173)
`er end by the spacer block, which houses the intermediate rods and provides a working space
`for access to the plunger system. Most units being installed in the oil field are of the horizontal
`configuration, which minimizes contamination of the crankcase oil with leakage from the fluid
`end. Vertical installations are still found, however, particularly with oil as the pumped fluid or
`when space is at a premium, as in townsite leases.
`Multiplex pumps applied to hydraulic pumping usually have stroke lengths from 2 to 7 in.
`and plunger diameters between 1 and 2½ in. The larger plungers provide higher flow rates but
`are generally rated at lower maximum pressure because of crankshaft loading limitations. The
`larger plungers provide higher flow rates, but are generally rated at lower maximum pressure
`because of crankshaft loading limitations. The normal maximum rating of multiplexes for con(cid:173)
`tinuous duty in hydraulic pumping applications is 5,000 psi, with lower ratings for the larger
`plungers, but applications above 4,000 psi are uncommon. Multiplex pumps are run at low
`speed to minimize vibration and wear and to avoid dynamic problems with the spring-loaded
`intake and discharge valves. Most applications fall between 200 and 450 rev/min, and because
`this is below the speeds of gas engines or electric motors, some form of speed reduction is
`usually required. Belt drives are found on some units, although gear reduction is more common
`while gear-reduction units are integral to some multiplexes and separate on others. A variety of
`reduction ratios are offered for each series of pumps. Because a positive-displacement pump
`has an essentially constant discharge flow rate for a given prime-mover speed, bypass of ex(cid:173)
`cess fluid normally is used to match a particular pressure and flow demand. Another option
`that has been used successfully is to drive the multiplex pump through a four-speed transmis(cid:173)
`sion, which greatly enhances the flexibility of the system. This allows much closer tailoring of
`the triplex output to the demand, thereby pumping at reduced speed when needed, which also
`tends to increase the life of such components as the packing and valving.
`
`GDI Ex. 1008
`Page 73 of 91
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`ASSSSS
`
`Chapter 14—Hydraulic Pumping in Oil Wells
`
`&>
`
`Bottomhole assembly
`with safety valve
`actuated by high-pressure
`powerfluid
`
`Fig. 14.16—Safety valve.
`
`PROD+PWRFLD IV-739 GDI Ex. 1008 ©
`
`Each plunger pumps individually from a common intake manifold into a common dis-
`charge, and because discharge occurs only on the upstroke,
`there is some pulsation, for which
`pulsation dampers are commonly used.
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`Page 74 of 91
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`IV-740
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`Petroleum Engineering Handbook- Vol. IV
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`\.>
`
`Fig. 14.17-Horizontal plunger pumps.
`
`Two types of plunger systems are in common use. For oi l service, a simple and effective
`plunger-and-liner system is used that consists of a closely fitted metallic plunger inside a metal(cid:173)
`lic liner. Sprayed metal coatings or other hard-facing means are often used to extend the life of
`the plunger and liner. When pumping water, the metal-to-metal system is not practical because
`the fit would have to be extremely close to keep leakage to an acceptable level. Galling and
`scoring are problems with close fits and the low lubricity of water, and to solve this problem,
`spring-loaded packing systems are used that do not require adjusting. The advent of high(cid:173)
`strength aramid fibers for packing, in conjunction with other compounds to improve the
`friction characteristics, has resulted in a pronounced improvement in the ability of the pump to
`handle high-pressure water for extended periods of time. Water still presents a more severe
`challenge than oil, however, and water systems show much better life if operated at or below
`3,500 psi.
`Suction conditions are importai1t to multiplex operation. Friction losses in piping, fluid end
`porting, and across the suction valving reduce the pressure available to fi ll the pumping cham(cid:173)
`ber on the plunger downstroke, and if these losses are sufficiently great, cavitation may result.
`When pumping oil with dissolved gas, the reduction in pressure liberates free gas and causes
`knocking, so it is necessa1y to have a positive head on the suction side to overcome the fric(cid:173)
`tion losses. In addition, another phenomenon known as "acceleration head" must be considered.
`The flow in the suction piping must accelerate and decelerate a number of times for each
`crankshaft revolution. For the fluid (which has inertia) to follow the acceleration, energy must
`be supplied, which is then returned to the fluid on deceleration. The energy supplied during
`acceleration comes from a reduction in the pressure in the fluid, and if this drops too low,
`cavitation or gas liberation will result. The minimum suction head for the multiplex pump is
`then the sum of the friction losses and the acceleration head. Although the pump can draw a
`vacuum, this will flash gas and may tend to suck air across the valve or plunger packing. Man(cid:173)
`ufacturers of multiplex pumps recommend appropriate suction charging pressures for their
`products, but it is worth noting that long, small-diameter suction lines increase the acceleration
`
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`Page 75 of 91
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`Chapter 14-Hydraulic Pumping in Oil Wells
`
`IV-741
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`head loss and friction loss. It is therefore recommended for suction lines to be short and of
`large diameter, with no high spots to trap air or gas. Suction stabilizers or pulsation dampeners
`that tend to absorb the pulsations from the pump also reduce acceleration head, and users are
`encouraged to follow good piping practices in the installation of surface pumps.
`In many cases, sufficient hydrostatic head is not available to provide the necessary suction
`pressure, and charge pumps are used to overcome this problem. Positive displacement pumps
`of the vane or crescent-gear type driven from the triplex have been used extensively, but they
`require a pressure-control valve to bypass excess fluid and match the multiplex displacement.
`Where electric power is available, centrifugal charge pumps have given excellent service. Cen(cid:173)
`trifugal pumps generally need to run at speeds considerably above the multiplex speed, and so
`driving them from the multiplex presents problems, particularly with a gas engine drive where
`prime-mover speed variations cause significant variations in the charge-pump output pressure.
`While good charging pressures are necessary to ensure proper loading and smooth opera(cid:173)
`tion, there are problems associated with very high charge pressures. These add to the
`crankshaft loading, and for charge pressures above about 250 psi, it is advisable to derate the
`maximum discharge pressure by one third of the charge pressure. High charge pressures also
`can adversely affect the lubrication of bearings, particularly in the crosshead wristpin. In addi(cid:173)
`tion, the mechanical efficiency of multiplex pumps is some 3 to 5% lower on the suction side
`compared to the discharge side. 22 Consequently, the combination of a charge pump and multi(cid:173)
`plex pump is most efficient with low charging pressures and a high boost by the multiplex
`pump. The charging pressure should therefore be limited to that necessary to give complete
`filling of the multiplex pump with a moderate safety allowance for variations in the system
`parameters.
`In some cases, it is desirable to inject corrosion inhibitors or lubricants into the multiplex
`suction, and fresh water is sometimes injected to dissolve high salt concentrations. In severe
`pumping applications with low-lubricity fluids, lubricating oil is sometimes injected or dripped
`onto the plungers in the spacer block area to improve plunger life. Injection pumps are often
`driven from the multiplex drive for these applications. A troubleshooting guide for multiplex
`pumps is given in Table 14.8.
`
`14.5.2 Fluid Controls. Various types of valves are used to regulate and to distribute the power-
`fluid supply to one or more wellheads. Common to all free-pump systems is a four-way valve
`or wellhead control valve, which is mounted at the wellhead, as shown in Fig. 14.18. Its func-
`tion is to provide for different modes of operation by shifting it to different positions. To
`circulate the pump into the hole, as shown in Fig. 14.15, power fluid is directed down the
`main tubing str