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ОПРЕДЕЛЕНИЕ СКОРОСТИ ИЗМЕНЕНИЯ ЧИСЛА ОБОРОТОВ ВАЛА НАСОСНОГО АГРЕГАТА ПРИ РЕГУЛИРОВАНИИ РЕЖИМА ПЕРЕКАЧКИ С УЧЕТОМ ОБЕСПЕЧЕНИЯ БЕЗОПАСНОСТИ ЭКСПЛУАТАЦИИ МАГИСТРАЛЬНОГО НЕФТЕПРОВОДА
DETERMINATION OF RATE OF PUMP SHAFT ROTATION SPEED CHANGE OF THE PUMPING AGGREGATE AT REGULATING THE PUMPING MODE TAKING INTO ACCOUNT THE SAFETY OF THE MAIN PIPELINE OPERATION
З. Х. Павлова
Z. Kh. Рavlova
Уфимский государственный нефтяной технический университет», г. Уфа
Ключевые слова: магистральный нефтепровод, магистральный насосный агрегат, технологический режим перекачки, регулирование режима перекачки, число оборотов вала насоса, давление в нефтепроводе, прочность труб, безопасность нефтепровода Key words: main oil pipeline, main line pumping unit, pumping process conditions, pumping mode control, pump shaft rotational speed, oil pipeline pressure, pipe durability, oil pipeline safety
Organizations operating the main pipelines work out their process conditions. Process conditions should provide for the required oil pumping rate with minimum running costs.
In the process of the main oil pipeline operation the required capacity may vary due to different reasons. In such cases oil pumping mode is regulated. The on-line pumping mode
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control can be carried out by turning the oil pump station main line pumps on and off and chocking oil flow. Chocking does not meet one of the basic process requirements - ensuring pumping with minimum running costs. The main line pump turning on and off leads to sharp changes in operating pressure and oil flow velocity and, subsequently, pressure surge in the pipeline [1^3]. The pressure surge is summed up with oil pipeline operating pressure, and the total pressure reaches significant values resulting in high stresses. High stress reduces pipe lifetime and the main line pump safety [4^6]. The pressure surge occurring in long operated oil pipelines with reduced pipe bearing capacity can cause accidents [4, 7].
The smoothest and economic pump operation control is provided by the main line pumps with variable-frequency electric drives. In conditions of the use of a variable-frequency electric drive, the pump shaft rotational speed can be smoothly changed by regulating the frequency converter outlet supply voltage frequency. The pump shaft rotational speed variation
Q1 <1
leads to oil pumping rate variation in accordance with the expression -=-, where Q1
Q2 c2
is oil flow rate at a pump shaft speed of C 1 , Q2 is the same at a speed of << .
Let's consider the problem of oil flow rate variation from Q1 to Q2 with the pump shaft speed changing. In order to ensure piping integrity and the main line pump safety, the pump shaft rotational speed should be controlled with the maximum mechanical stresses in the pipe wall not exceeding the standard ones. Besides, considerable reduction of the above total maximum stresses should be provided. These conditions can be obtained by smooth oil flow rate changing from Q1 to Q2 within a certain period of time t with the use of a variable-frequency electric drive. Let's take this period of time t through flow rates Q1 and Q2, pipe cross-sectional area in clear Fclear = 0,25 -n •D2int and the pipeline section length £0 , where
uniform flow rate change from Q l to Q 2 occurs, as
n-D2, •10
t = —;- 0 x , (1)
2 •( Q1C + Q2C )
where Dnt is the internal oil pipe diameter, m; Q1 , Q2 are flow rates (capacities), m3/sec.
During pumping mode analysis it is necessary to take into account the fact that within one and the same process section operating modes of all the main line pumps of oil pumping stations are interdependent. Process sections are separated from each other by tanks.
Taking into account the above statements and papers [3, 8, 9], based on the use of the discharge head balance equation for determination of the rotational speed C2 of the controlled main line pump shaft with a variable-frequency electric drive, providing for capacity Q 2, we have
< = CN • F0,5, (2)
where F = (QfcTK + Q^^ - a2 + h0 -az); (3)
S ac
K = ; (4)
Q 2 is flow rate (capacity), m3/h;
Np Np b
ub
a2 = ab + S ai ; Ь2 = Sb + f b 2-m +S bc
1 1 (Nb ) 1
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NP is the number of the main line pumps running; Nc is the number of the controlled main line pumps running; Nb is the number of paralleled booster pumps; ac, ai, ab , bc, b , bb are head-capacity characteristics of the main line pumps (controlled and uncontrolled) and those of booster pumps [8, 9]; a>N is the pump shaft nominal speed; h is the residual discharge head at the end of the process section; Az = z1 - z2, z1, z2 is the geodetic mark of the beginning and end of the process section; L is length of the oil pipeline process section; P, m are coefficients dependent on the fluid flow pattern [10]; v is kinematic viscosity of oil.
Analysis of the analytical expressions (2)^(4) shows that increase in the process section length L and the oil viscosity v , as well as the oil pipeline diameter Dn reduction lead to the necessity to increase the controlled pump shaft speed a)2 in order to ensure the required pumping rate Q 2 .
It should be noted that the exact calculation of the K parameter by (4) proves to be rather complicated due to some possible changes in oil properties and uncertainty of head losses on the local resistances of the calculated oil pipeline section. In order to simplify and clarify the calculation, the actual data on oil pumping characteristics can be used. Thus, for example, the K parameter is recommended to be determined by the following formula, using the actual data for the calculated section
H, + - h0
K -0-, (5)
02-m ' v 7
1С
where Hd is the pump station discharge head, determined on the basis of the instrument data in the process of pumping with a capacity QC.
The formula (5) can be used for K determination in cases when changes in flow rate from Q1 to Q2do not result in P and m coefficient changes.
In papers [1,8] it is pointed out that sharp changes in the pipeline pressure result in occurrence of pipe wall local radial displacements, additional circumferential and bending stresses. Based on the studies of oil pipe wall stress and strain state in conditions of interior pressure variation in the section l0, we have obtained calculating formulae for determination of circumferential and bending stresses. Thus, circumferential stresses and bending (longitudinal) stresses occur simultaneously in the design section. In this context, in accordance with the requirements [11], equivalent stresses were also determined. The maximum equivalent stresses exceed circumferential and bending stresses. The maximum equivalent stress values are determined by the following formula
p •R ~
~S
<7 . =--<7 . (6)
equiv g eqrnv ' \ /
where
2 f \2 ^P 1 , 4 5824 I bP 1 + 2,9051- AP
1 + | + 4,5824• P
p • a
0
P • a0
(7)
where p is pipeline pressure at a flow rate Q1 ;
Ap is pressure increase value during flow rate variation from Q1 to Q2, taking into account change in operating pressure due to flow rate variation and local increase in pressure resulting from flow velocity variation, determined in accordance with the recommendations in papers [1, 8];
8 is pipe wall thickness;
R is pipe wall mids [in Russian]. urface radius;
0,5
и
equivm
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a0 is a nondimensional parameter, determined by the following formula
1,2854 ■£ „
( RS
In accordance with the requirements [11] and in order to ensure pipe durability, the following condition should be observed:
o < R C, (9)
equiv 2 ' v '
wher RC is characteristic strength of pipe metal, equal to yield point oT .
Analysis of the analytical expressions (6)^(8) shows that increase in length £0 leads to equivalent stress crequiv reduction. Besides, theses expressions allow determination of such a £0 value, which will ensure compliance with the regulatory pipe durability requirements in conditions of pumping rate variation.
Based on the regulations [11] and using the analytical expressions given above, length l0 can be determined by applying the following condition:
where
10 > 0,7780-a0c (R-S)0
1,4526-AP |1 +AP
B
1 +
1 + -
2,1718-B
\ 2 1 + Ï )
B = (RHГ-|1I ; R2H =
ap v . rh _ Rh -s
p-R
(10)
(11)
Alternating high stresses in the main line pump pipe metal including those at standard level, are known to shorten the oil pipeline life [5]. With this statement taken into account, the length £0 value should be determined, where an increase in stress resulting from pumping rate variation from Q1 to Q 2 will not exceed the set value o which is quite reliable for the main line pump reliability and safety. Without considering a local increase in pressure
. . . . . . . P2 R -c
resulting from oil flow variation, the maximum equivalent stress value will be —;— o
S
where crc is a nondimensional parameter determined by (7), replacing Ap with Apn,
equiv .m J 1. 1 O
where Apo is a change in operating pressure resulting from pumping rate variation.
Then based on the above condition of stress increase restriction by C , using (6) and (7),
we have
p-R (cr .
S \ equiv.m
-c _c,
(12)
In accordance with the expression (7), <Jequ- m u O parameters contain a0 .
eqUlV .m equiv.m 0
Admitting the <O value from the main line pump safety and reliability conditions (12), the a 0 parameter can be determined, and the £ 0 length value can be calculated by (8). Then time t is determined by (1) with preset Q1 to Q2, and the controlled pump shaft rotational
speed 0)2 is calculated by (2) for the required pumping rate Q2 in conditions of ensuring pipe duration and the main oil pipeline operation safety.
ao _
0.5
a0c =
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As an example, let's consider an oil pipeline 0 1020 *11 mm process section from an oil pump station, where NM 7000-210 main line pumps and NPV 3600-90 booster pumps are located. The section length is 150 km, Az = -50 m, h = 40 m. Vertical lift performance of the pumps and oil properties are taken in accordance with [10]. Pipe material is 17G1S steel with a yield point <Y =366 MPa. With two running main line pumps (connected in series) and two booster pumps (paralleled), pumping rate is 5000 m3/h. Based on the calculating formulae presented in papers [8, 10] with the use of formulae (2)^(4), it was determined that to get pumping rate Q 2 =6000 m3/h additional smooth activation of one main line pump with
a variable-frequency electric drive is required providing that its shaft speed (02 = 0,75- ()N . For significant reduction of pressure increase due to variation of pumping rate from 5000 m3/h to 6000 m3/h (or from 1.3889 m3/sec to 1.6667 m3/sec) and oil flow velocity from 1,78 m/sec to 2,13 m/sec, in accordance with the recommendations in paper [8] and taking into
account calculating formulae (8) и (10), length i 0 can be admitted to be 100m. The controlled pump rotational speed and oil pumping rate should be changed within 51 sec in order to ensure the main oil pipeline safety. The maximum pipe wall equivalent stress is 243.1 MPa which is much less than the pipe metal yield point.
Conclusions
1. Based on the analysis of pumping rate change process with the use of the main line pumps with a variable-frequency electric drive, a calculation formula for determination of pumping rate variation time period is proposed following piping durability and oil pipeline safety requirements.
2. Analytic dependences were obtained to determine a tolerance extent for the oil pipeline section length where pipeline oil flow velocity change occurs. Pipe wall stresses will not exceed standard values in this case. A calculating formula is proposed allowing establishment of such oil pumping mode change parameters when local increase in pressure and stress resulting from fluid flow velocity variation will not exceed the value preset to ensure the oil pipeline safety.
3. Based on the analytic dependencies obtained, it is possible to determine the rotational speed of the main line pump shaft with a variable-frequency electric drive along with pumping mode change taking into account pipe durability and the main oil pipeline operation safety.
References
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8. Шабанов В. А., Алексеев В. Ю., Павлова З. Х. Обеспечение бесперебойной работы частотно-регулируемых электроприводов магистральных насосов и технологического режима перекачки при кратковременных нарушениях электроснабжения: монография. Уфа: Нефтегазовое дело. 2012. - 171 с.
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Сведения об авторе
Павлова Зухра Хасановна, к. т. н, доцент кафедры «Электротехника и электрооборудование предприятий», декан факультета автоматизации технологических процессов, Уфимский государственный нефтяной технический университет», г. Уфа, тел. 8(347)2420851, e-mail: [email protected]
Pavlova Z. Kh., Candidate of Science in Engineering, associate professor of the chair «Electrical Engineering and Electrical Equipment»; dean of the Faculty of Automation of Industrial Processes; Ufa, phone: 8(347)2420851, e-mail: [email protected]
