Study on Key Factors Influencing the ROP Improvement of PDC Bits
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摘要:
为了在钻井工程中发挥出PDC钻头的最大功效,通过理论分析、室内试验、案例分析、现场试验等,探讨了高钻压、高转速等钻井参数强化对PDC钻头钻速和磨损的影响规律,同时分析了PDC钻头的磨损机理与过早失效主因。研究结果表明:1)钻压是影响PDC钻头机械钻速的直接和首选因素,当钻头处于高效破岩状态时,无论钻遇一般地层还是硬岩地层,钻压与机械钻速均应呈线性关系;钻遇均质硬岩地层时,建议将200 kN以上高钻压纳入PDC钻头的常规应用参数;2)提高转速可实现钻井提速,虽然高转速会加剧PDC钻头的磨损,但目前切削齿的质量足以满足PDC钻头在高转速(400~500 r/min)下长时间钻进多数地层的需求;3)布齿密度对钻头机械钻速有影响,但并非直接因素,只要“吃得进去,切得下来,排得及时”三者建立动态平衡,即便是高布齿密度PDC钻头也可以实现优快钻进;4)PDC钻头破岩效率越高,钻头磨损会越小,如提高钻压,会增大切削齿吃入深度、减少钻头磨损;5)动态冲击和低效破岩是造成PDC切削齿和钻头过早失效的主因,实现PDC钻头高效钻进的核心是提高破岩效率与抑制钻头振动。该研究结果对PDC钻头合理使用与钻井提速技术创新具有参考意义。
Abstract:For the maximization of the efficacy of the polycrystalline diamond compact (PDC) bits in drilling engineering, comprehensive research, including theoretical analysis, laboratory test, case study, and on-site trials, was conducted to investigate how a high weight-on-bit (WOB), a high rotary speed, and other optimized drilling parameters work on the rate of penetration (ROP) and the wear of a PDC bit. Furthermore, the wear mechanism of the PDC bit and the primary cause of the premature failure of the bit were analyzed. The results indicated that: 1) The ROP of the PDC bit was directly and primarily affected by the WOB. When the bit was in an efficient rock-breaking state, the WOB was invariably in a linear relationship with the ROP whether the formation encountered was a conventional one or a hard rock formation. Adding a high WOB over 200 kN into the normal pressurization range of the PDC bit was recommended if the formation encountered was a homogeneous hard rock formation. 2) ROP improvement could be achieved by enhancing the rotary speed. Although the wear of the PDC bit could be aggravated by a high rotary speed, the requirement on a PDC bit to penetrate most formations for a long time at a high rotary speed (400–500 r/min) could be readily met by the quality of the currently available PDC cutter. 3) The ROP of the bit was also affected by cutter density, but not in a direct manner. As long as a dynamic balance among “capabilities to bite into the formation, cut the rock, and evacuate the cuttings in time” was reached, the optimized fast drilling could be achieved even by a PDC bit with a high cutter density. 4) The wear of the PDC bit was less severe under the higher rock-breaking efficiency of the bit. The WOB could be enhanced to improve the ROP and reduce bit wear. 5) Dynamic impact and inefficient rock-breaking were considered the primary causes of the premature failure of the PDC cutter and bit. The key for the PDC bit to achieve efficient penetration was improving rock-breaking efficiency and restraining bit vibration. The above results could be used as a reference for the proper utilization of PDC bits and the innovation of ROP improvement technologies.
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川渝地区页岩层经历了强烈的后期改造,地质条件相对复杂,页岩分布不稳定,呈现较强的各向异性特征。对于页岩气的勘探开发,井壁取心技术是关键技术之一,页岩气水平井的水平段长达1 000~2 000 m,采用常规钻杆、连续油管难以将取心器准确下至取心位置,且钻井完井工作难度大、耗时长、费用高[1-2]。针对川渝地区页岩气水平井长水平段取心困难的问题,张宇奇[3]将井下爬行器与旋转式井壁取心器相结合,设计了一种具备爬行、定位、推靠、取心、储样和解卡等功能的旋转式井壁取心器,可完成水平井水平段、大位移定向井斜井段的取心作业;张朝界等人[4]用Solidworks软件模拟实际工况,建立了页岩气水平井和取心器的三维模型,利用ADAMS虚拟样机仿真技术,对取心器的爬行能力、过弯能力、负载能力和越障能力进行了模拟分析,结果表明均满足设计要求。
爬行机构作为旋转式井壁取心器的直接驱动装置,其性能决定了取心器能否正常完成井下取心工作。1994年,J. Hallundbæk首先设计了Welltec轮式爬行器[5],Sondex公司对轮式爬行器进行了改进,采用了2个扶正机构[6];D. Bloom等人[7]研发了Maxtrac伸缩式爬行器,M. Buyers等人[8]对其进行了改进,可以蠕动前进。2001年,沈阳工业大学研制了管道爬行器;高进伟等人[9]根据平行四边形原理设计了爬行及定心装置,解决了爬行器在井下的轴向居中问题;周劲辉等人[10]研制了水平井自扶正式电缆爬行器;唐德威等人[11]研制了井下电机驱动爬行器。适用于水平井的爬行器以伸缩式爬行器和轮式爬行器为主,伸缩式爬行器的负载大,但爬行速度较慢;轮式爬行器的爬行速度快,但牵引力较小,仅能完成测井工具的运输,无法携带大段岩心,不适用于川渝地区页岩气水平井长水平段的取心工作。为此,笔者对传统爬行机构进行改进,采用行星齿轮、锥齿轮组合的传动方式,利用正交试验分析方法,分析各因素对支撑臂伸出速度和支撑臂推靠力的影响程度,并对主要结构尺寸进行了优化,降低了支撑臂所需要的推靠力,提高了支撑臂的伸出速度。
1. 爬行机构传动方案设计
取心器的爬行轮要求有足够的扭矩和正压力,以克服摩擦阻力、井下水平段及造斜段电缆拖拽力和井下流体阻力,而爬行轮的扭矩需要液压或电机来提供。由于整体尺寸的限制,要求爬行轮的转动速度及转矩较高,且因井下温升问题无法使用液压驱动来提供动力,只能用电机驱动爬行轮转动。为了满足取心直径要求,所选择电机的直径不能太大;考虑整个取心器系统需要地面提供电力,要求地面采用高压输电方式进行供电,相应地需要选择高压电机;由于尺寸控制,电机转速越高,电机尺寸越小。综合考虑,选择特制高速电机。
取心器的前进动力由爬行轮提供,需要选择多种传动方式来实现电机与爬行轮之间的传动。行星齿轮减速器具有同轴向输出扭矩、轴向尺寸小和传动比大等特点,而且体积微小,可适用于精密仪器、电动装置、操作机构和取心器系统等设备;蜗轮蜗杆传动结构紧凑,单级传动比大,工作较稳定,但安装精度要求高,不适合用于爬行轮传动;带传动适用于高速传动,且安装时需要一定预紧力,无法在爬行轮传动过程中使用。因此,选择行星齿轮作为主要动力传动,搭配可以改变传动方向的锥齿轮,驱动电机的动力经过行星齿轮减速器、锥齿轮、链传动和行星爬行轮到达爬行轮,从而实现取心器的爬行功能。设计的爬行机构传动方案见图1。
2. 爬行机构优化设计
爬行臂作为爬行机构的主要部件,一方面可以作为传动机架,把锥齿轮的动力通过链传动传递到爬行轮上;另一方面,爬行臂作为伸出部分,其末端装配爬行轮,与支撑臂相互配合,完成爬行轮的压紧工作。支撑臂作为支撑调节机构,对运动状态进行微调,达到取心器所需要的预压力。所以,二者作为爬行机构的主要部件,其结构尺寸和结构强度对整个机构的性能影响非常大。
2.1 爬行臂和支撑臂受力分析
爬行臂和支撑臂的受力如图2所示(O为爬行臂铰接点,C为爬行轮中心)。
根据几何关系,可得[12]:
sinα=D−d2a (1) sinβ=D−d+2bsinα+2e2c (2) 式中:D为井筒直径,mm;a,b为爬行臂CA段和OC段的长度,爬行臂OA的长度为a+b,mm;c为支撑臂AB的长度,mm;d为爬行轮直径,mm;e为支撑臂铰接点与轴线的偏心距,mm;α为爬行臂转角,(°);β为支撑臂转角,(°)
对爬行臂和支撑臂进行受力分析,可得平衡方程:
{F0sinα+FBcosβ=FNF0cosα+FBsinβ=0 (3) 式中:F0为爬行臂正压力,N;FB为支撑臂B点推靠力,N;FN为爬行轮所受正压力,N。
由此,得到支撑臂和爬行臂的力矩公式为:
{FNacosα−(F0sinα+FAcosα)(a+b)sinα=0FAcsinαsinβ−FAccosαcosβ=0 (4) 式中:FA为A点的推靠力,N。
由于支撑臂上A点和B点的力在各自方向上的分力大小相同、方向相反,联立式(3)和式(4)可得:
{F0cosα=FAsinα=FBsinβFAcosα+F0sinα=FNFNacosα−(a+b)FAsinα(cosαtanβ+sinα)=0 (5) 则爬行轮所受摩擦力fP为:
fP=μFN=μ(a+b)(tanα+tanβ)aFBsinβ (6) 式中:μ为爬行器与井壁的摩擦系数,理论上可取0.5;fP为爬行轮所受摩擦力,N。
取心器所需要的总推进力为6 000 N,爬行轮设计为2组,每组有3个爬行轮,则单个爬行轮所要达到的正压力为1 000 N。
2.2 爬行臂优化分析
在液压缸推力的作用下滑块移动,推动支撑臂伸出,爬行臂绕O点旋转,带动爬行轮压靠在井壁上(见图3)。爬行机构的基本性能参数是爬行轮的正压力及其工作效率。工作效率主要取决于支撑臂的伸出速度,支撑臂的伸出速度由滑块位移决定,爬行臂长度a+b、支撑臂长度c、爬行臂转角α和偏心距e等因素都会对其产生影响。将这4个影响因素确定为优化变量,建立爬行机构的优化设计函数。
根据几何关系,可得:
{s=sA+√c2−[(a+b)sinα+e]2s=(a+b)−(a+b)cosα+ccosψe=csinψ (7) 式中:sA为滑块位移,mm;s为滑块右死点距铰接点A的水平距离,mm;
ψ 为支撑臂初始转角,(°)。化简式(7),可得滑块位移sA的计算式:
sA=ccosψ−√c2−[(a+b)sinα+e]2+(a+b)(1−cosα) (8) 式(8)对时间求导,可得支撑臂伸出速度vA的计算式:
vA=(a+b)sinα−e+[(a+b)sinα+e](a+b)cosα√c2−[(a+b)sinα+e]2 (9) 式中:vA为支撑臂的伸出速度,mm/s
根据式(6),可得爬行轮所受正压力FN为:
FN=(a+b)(tanα+tanβ)aFBsinβ (10) 根据几何关系及式(1)、式(2),可得:
sinβ=(a+b)sinα+ec (11) tanβ=(a+b)sinα+e√c2−[(a+b)sinα+e]2 (12) 将式(11)、式(12)代入式(10),可得到支撑臂推靠力FB:
FB=acFN(a+b)(tanα+(a+b)sinα+e√c2−[(a+b)sinα+e]2)[(a+b)sinα+e] (13) 式(9)和式(13)即为爬行机构的多目标优化函数,利用正交试验分析法对其进行分析,可得到支撑臂及爬行臂有关参数的优化解。
影响支撑臂伸出速度和推靠力的因素包括爬行臂长度a+b、支撑臂长度c、爬行臂转角α和偏心距e。已知爬行轮直径为60 mm,取心器适用于ϕ200.0 mm的水平井,可以确定各影响因素的参数水平(见表1)。
表 1 正交试验各因素的水平Table 1. Factor level of the orthogonal test水平 因素 a/mm b/mm c/mm e/mm α/(°) 1 75 20 100 5 30 2 80 25 115 8 35 3 102 28 130 12 40 4 120 30 140 15 45 表 2 正交试验方案及结果Table 2. Plan and results of the orthogonal test序号 因素 vA/
(mm·s–1)FB/N a/mm b/mm c/mm e/mm α/(°) 1 75 20 100 5 30 43.034 1 476.282 2 75 25 115 8 35 49.923 948.763 3 75 28 130 12 40 54.784 760.183 4 75 30 140 15 45 59.831 613.171 5 80 20 115 12 45 59.442 546.541 6 80 25 100 15 40 53.611 401.823 7 80 28 140 5 35 57.392 1 244.544 8 80 30 130 8 30 47.480 1 326.465 9 102 20 130 15 35 55.684 817.832 10 102 25 140 12 30 52.055 1 222.991 11 102 28 100 8 45 102.036 29.503 12 102 30 115 5 40 80.807 473.042 13 120 20 140 8 40 82.741 673.195 14 120 25 130 5 45 98.571 404.759 15 120 28 115 15 30 60.058 582.237 16 120 30 100 12 35 78.109 143.874 为了确定上述各因素对试验指标的影响,将求解的指标进行极差计算,即可找出各因素的主次顺序及优化组合,结果见表3和表4。正交试验各指标的平均值用ki(i=1,2,3,4)表示,其中i表示每个变量的因素水平顺序,将各指标平均值进行极差处理。极差R表示目标量变化的最大范围,可以用来表征不同变量对指标值的影响程度。指标值越大,此变量对目标函数的影响程度越大,需要重点考虑;指标值越小,此变量对目标函数的影响越小,可优先满足其他指标后再进行考虑[13-14]。
表 3 支撑臂伸出速度极差分析结果Table 3. The extension speed range analysis of the supporting arm参数 不同因素水平对应的支撑臂伸出速度vA /(mm·s–1) a b c e α k1 51.799 60.225 69.197 69.951 50.657 k2 54.481 63.540 62.558 70.545 60.277 k3 72.645 68.568 64.130 61.098 67.985 k4 79.870 66.557 63.005 57.296 79.970 R 28.071 8.343 6.639 13.249 29.313 表 4 支撑臂推靠力极差分析结果Table 4. The push-the-bit force range analysis of the supporting arm参数 不同因素水平对应的支撑臂推靠力FB/N a b c e α k1 949.550 878.460 512.870 899.650 1 151.990 k2 879.325 744.580 637.650 744.480 788.750 k3 635.830 374.120 827.310 668.400 577.040 k4 451.000 638.910 938.470 603.750 398.490 R 498.550 504.340 425.600 295.900 753.500 根据多目标优化理论,对2个目标量vA和FB进行分析,得出各因素的影响程度:各因素对目标函数vA的影响程度从大到小的顺序为α,a,e,b和c;对目标函数FB的影响程度从大到小的顺序为α,b,a,c和e。比较2组目标函数的优化值,可首先确定α,a和b的优化解分别为45°,120 mm和30 mm。通过比较影响程度的大小,得到e的优化解为8 mm,根据爬行臂长度确定c的优化解为140 mm。
为了确定求得的优化解对爬行机构试验指标的影响,将优化解代入原目标函数,并与优化前各结构尺寸的试验指标进行对比,结果见表5。
表 5 优化前后试验指标对比Table 5. Comparison between test indicators before and after optimization优化前后 a/mm b/mm c/mm e/mm α/(°) vA/(mm·s–1) FB/N 优化前 75 20 100 5 30 43.034 1 259.222 80 25 115 8 35 52.829 893.596 102 28 130 12 40 72.393 554.946 120 30 140 15 45 92.284 339.869 优化后 120 30 140 8 45 99.060 408.230 从表5可以看出,根据正交试验结果优选出的结构尺寸可以降低支撑臂所需推靠力,提高支撑臂伸出速度,说明可以使用正交试验方法优化爬行机构的结构尺寸,优化结果满足要求。
2.3 爬行机构结构设计
根据爬行臂和支撑臂的优化设计结果,对爬行机构的各零部件进行设计、选型、强度校核和刚度校核,使用Solidworks软件对其进行建模和虚拟装配,得到了爬行机构的三维模型,如图4所示。
3. 结 论
1)根据页岩气井旋转式井壁取心器的工作要求,设计了一种由行星齿轮、锥齿轮组合传动的新型爬行机构,能够带动整个取心器行进。
2)根据机械动力学原理,建立了爬行机构正压力、支撑臂伸出速度、支撑臂推靠力与爬行臂及支撑臂结构尺寸的函数方程。
3)爬行臂转角对支撑臂伸出速度和推靠力影响最大。爬行臂转角优化后,可以降低支撑臂所需推靠力,提高支撑臂伸出速度,优化结果满足要求。
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表 1 VTL试验参数
Table 1 Vertical turning lathe (VTL) test parameters
试验
编号每圈吃入
深度/mm总的行进
距离/m切削深度/
mm线速度/
(m·min−1)#1 0.5 68 097 60 100 #2 1.0 34 049 #3 1.5 22 699 #4 2.0 17 024 #5 2.5 13 619 #6 3.0 11 350 #7 1.0 34 049 60 20 #8 34 049 60 #9 34 049 100 #10 34 049 140 表 2 高速螺杆与常规螺杆参数对比
Table 2 Parameter comparison between high-speed motor and conventional motor
螺杆类型 钻压/kN 工作排量/
(L·min−1)输出扭矩/(N·m) 顶驱转速 /
(r·min−1)钻头转速 /
(r·min−1)ϕ172.0 mm高速螺杆 60~100 2200 8869 60~80 380~400 ϕ172.0 mm常规螺杆 60~150 2200 12750 60~80 220~240 表 3 玛南风城组不同钻具组合的钻井指标
Table 3 Drilling performances of various bottom-hole assemblies in Fengcheng Formation on southern slope of Mahu Sag
试验井 钻头 井下动力钻具 单趟平均进尺/m 平均机械钻速/(m·h−1) 井型 完钻时间 JL53井 牙轮钻头、PDC钻头、复合钻头 <50 <1.3 直井 2020年 JL56井 异形齿PDC钻头 常规螺杆 88 2.0 直井 2020年 MH48井 孕镶钻头 涡轮 193 1.8 直井 2020年 MN520井 PDC钻头 旋导 86 1.2 水平井造斜段 2021年 PDC钻头 高速螺杆 578 4.8 水平井水平段 MN272井 PDC钻头 高速螺杆 1008 8.2 水平井水平段 2022年 表 4 胜利油田常规钻井参数与强化钻井参数对比
Table 4 Comparison of conventional and enhanced drilling parameters in Shengli Oilfield
钻井参数类型 钻压/kN 顶驱转速 /(r·min−1) 排量/(L·s−1) 泵压/MPa 提速工具 常规钻井参数 40~80 70 40 15 螺杆 强化钻井参数 100~120 70~80 >70 >20 大扭矩螺杆 表 5 美国FORGE 78B-32井TKC83型PDC钻头钻井指标
Table 5 Drilling data of TKC83 PDC bit in FORGE Well 78B-32
趟钻数 钻头直径/mm 入井井深/m 进尺/m 平均机械钻速/(m·h−1) 钻压/kN 顶驱转速/(r·min−1) 排量/(L·s−1) 钻遇岩性 7 269.9 1 112.8 643.1 20.4 295 40.0 51.7 花岗闪长岩 9 269.9 1 774.2 267.9 22.3 295 50.0 50.5 花岗闪长岩 13 269.9 2 055.0 265.5 21.2 295 45.0 52.4 花岗闪长岩 14 269.9 2 320.5 270.4 22.5 295 50.0 52.4 花岗闪长岩 -
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