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PDC钻头钻井提速关键影响因素研究

高德利, 刘维, 万绪新, 郭勇

高德利,刘维,万绪新,等. PDC钻头钻井提速关键影响因素研究[J]. 石油钻探技术,2023, 51(4):20-34. DOI: 10.11911/syztjs.2023022
引用本文: 高德利,刘维,万绪新,等. PDC钻头钻井提速关键影响因素研究[J]. 石油钻探技术,2023, 51(4):20-34. DOI: 10.11911/syztjs.2023022
GAO Deli, LIU Wei, WAN Xuxin, et al. Study on key factors influencing the ROP improvement of PDC bits [J]. Petroleum Drilling Techniques,2023, 51(4):20-34. DOI: 10.11911/syztjs.2023022
Citation: GAO Deli, LIU Wei, WAN Xuxin, et al. Study on key factors influencing the ROP improvement of PDC bits [J]. Petroleum Drilling Techniques,2023, 51(4):20-34. DOI: 10.11911/syztjs.2023022

PDC钻头钻井提速关键影响因素研究

基金项目: 国家自然科学基金重点项目“复杂结构‘井工厂’立体设计建设基础研究”(编号:52234002)、国家自然科学基金创新研究群体项目“复杂油气井钻井与完井基础研究”(编号:51821092)、中国石油大学(北京)科研启动基金项目“高效钻头的研究”(编号:ZX20190065)联合资助
详细信息
    作者简介:

    高德利(1958—),男,山东禹城人,1982年毕业于华东石油学院钻井工程专业,1984年获西南石油学院石油矿场机械专业硕士学位,1990年获石油大学油气田开发工程专业博士学位,教授,中国科学院院士,长期从事复杂油气井工程领域的科学研究与实践。系本刊编委。E-mail: gaodeli@cup.edu.cn。

  • 中图分类号: TE21

Study on Key Factors Influencing the ROP Improvement of PDC Bits

  • 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.

  • 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。

    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

    图  1  爬行机构传动设计方案
    Figure  1.  Transmission design of the crawling mechanism

    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。

    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)。

    图  2  爬行臂和支撑臂力学分析
    Figure  2.  Mechanical analysis of the crawling arm and supporting arm

    根据几何关系,可得[12]

    sinα=Dd2a (1)
    sinβ=Dd+2bsinα+2e2c (2)

    式中:D为井筒直径,mm;ab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=0FAcsinαsinβFAccosαcosβ=0 (4)

    式中:FAA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。

    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。

    图  3  爬行机构运动简图
    Figure  3.  Kinematic sketch of the crawling mechanism

    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=ccosψc2[(a+b)sinα+e]2+(a+b)(1cosα) (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)sinαe+[(a+b)sinα+e](a+b)cosαc2[(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+ec2[(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+ec2[(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/mmb/mmc/mme/mmα/(°)
    1 7520100530
    2 8025115835
    31022813012 40
    41203014015 45
    下载: 导出CSV 
    | 显示表格

    根据表1设计的正交试验方案,共进行16次试验,结果见表2

    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/mmb/mmc/mme/mmα/(°)
    1 752010053043.0341 476.282
    2 752511583549.923948.763
    3 752813012 4054.784760.183
    4 753014015 4559.831613.171
    5 802011512 4559.442546.541
    6 8025100154053.611401.823
    7 802814053557.3921 244.544
    8 803013083047.4801 326.465
    91022013015 3555.684817.832
    101022514012 3052.0551 222.991
    1110228100845102.036 29.503
    121023011554080.807473.042
    131202014084082.741673.195
    141202513054598.571404.759
    151202811515 3060.058582.237
    161203010012 3578.109143.874
    下载: 导出CSV 
    | 显示表格

    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表4。正交试验各指标的平均值用kii=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
    abceα
    k151.79960.22569.19769.95150.657
    k254.48163.54062.55870.54560.277
    k372.64568.56864.13061.09867.985
    k479.87066.55763.00557.29679.970
    R28.071 8.343 6.63913.24929.313
    下载: 导出CSV 
    | 显示表格
    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
    abceα
    k1949.550878.460512.870899.6501 151.990
    k2879.325744.580637.650744.480788.750
    k3635.830374.120827.310668.400577.040
    k4451.000638.910938.470603.750398.490
    R498.550504.340425.600295.900753.500
    下载: 导出CSV 
    | 显示表格

    根据多目标优化理论,对2个目标量vAFB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αaebc;对目标函数FB的影响程度从大到小的顺序为αbace。比较2组目标函数的优化值,可首先确定αab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/mmb/mmc/mme/mmα/(°)vA/(mm·s–1FB/N
    优化前 752010053043.0341 259.222
    802511583552.829893.596
    1022813012 4072.393554.946
    1203014015 4592.284339.869
    优化后1203014084599.060408.230
    下载: 导出CSV 
    | 显示表格

    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。

    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。

    图  4  爬行机构的三维模型
    Figure  4.  Three-dimensional model of the crawling mechanism

    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。

  • 图  1   钻头高效破岩时钻压与机械钻速的关系示意

    Figure  1.   Relationship between WOB and ROP during efficient rock-breaking of the bit

    图  2   岩性和齿形对钻压与机械钻速之间关系曲线的影响

    Figure  2.   Effects of lithology and cutter shape on relationship curve between WOB and ROP

    图  3   “异常”因素作用时钻压与机械钻速的关系示意

    Figure  3.   Relationship between WOB and ROP under influences of “abnormal” factors

    图  4   脱钴PDC切削齿的磨损面积与其行进距离的关系

    Figure  4.   Relationship between wear area and travel distance of leached PDC cutter

    图  5   相同进尺下PDC钻头吃入深度与切削齿行进距离的关系

    Figure  5.   Relationship between cut depth of PDC bit and travel distance of cutter under the same drilling footage

    图  6   相同进尺下PDC钻头吃入深度与切削齿磨损体积的对应关系

    Figure  6.   Relationship between the wear volume loss of PDC cutter and the cut depth under the same footage

    图  7   转速对PDC切削齿磨损体积的影响

    Figure  7.   Effect of rotary speed on wear volume of PDC cutter

    图  8   胜利油田罗家区块二开钻井指标

    Figure  8.   Drilling data from Luojia block in Shengli Oilfield

    图  9   NPD的耐磨性和抗冲击性测试示意

    Figure  9.   Wear resistance and impact resistance tests of nano-polycrystalline diamond (NPD)

    图  10   135°斧形齿钻遇花岗岩时发生冲击失效

    Figure  10.   Impact-induced failure of 135° axe-shaped teeth when encountering granite

    图  11   PDC钻头的破岩、耐用、稳定一体化综合评价体系示意

    Figure  11.   Comprehensive evaluation system integrating rock-breaking efficiency, durability, and stability of PDC bit

    图  12   PDC切削齿聚晶金刚石层的横截面

    Figure  12.   Cross-sections of polycrystalline diamond layer of PDC cutter

    图  13   脱钴和未脱钴PDC切削齿的磨损体积与行进距离的关系

    Figure  13.   Relationships between wear volumes and travel distances of leached and non-leached PDC cutters

    图  14   抗冲击性测试后的未脱钴PDC切削齿形貌

    Figure  14.   Morphology of non-leached PDC cutter after impact resistance test

    图  15   PDC切削齿的典型出井状况

    Figure  15.   Typical dull conditions of PDC cutters pulled out of hole

    图  16   X射线检测的PDC切削齿脱钴深度

    Figure  16.   Leached depth of PDC cutters detected by X-ray

    图  17   未脱钴PDC切削齿的出井形貌

    Figure  17.   Morphology of non-leached PDC cutters pulled out of hole

    图  18   与图7对应的切削齿磨口形貌

    Figure  18.   Wear scar morphology of PDC cutter corresponding to Fig.7

    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
    #10.568 09760100
    #21.034 049
    #31.522 699
    #42.017 024
    #52.513 619
    #63.011 350
    #71.034 0496020
    #834 04960
    #934 049100
    #1034 049140
    下载: 导出CSV

    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60~1002200886960~80380~400
    ϕ172.0 mm常规螺杆60~15022001275060~80220~240
    下载: 导出CSV

    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882.0直井2020年
    MH48井孕镶钻头涡轮1931.8直井2020年
    MN520井PDC钻头旋导861.2水平井造斜段2021年
    PDC钻头高速螺杆5784.8水平井水平段
    MN272井PDC钻头高速螺杆10088.2水平井水平段2022年
    下载: 导出CSV

    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40~80704015螺杆
    强化钻井参数100~12070~80>70>20大扭矩螺杆
    下载: 导出CSV

    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
    7269.91 112.8643.120.429540.051.7花岗闪长岩
    9269.91 774.2267.922.329550.050.5花岗闪长岩
    13269.92 055.0265.521.229545.052.4花岗闪长岩
    14269.92 320.5270.422.529550.052.4花岗闪长岩
    下载: 导出CSV
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出版历程
  • 收稿日期:  2022-12-04
  • 修回日期:  2023-01-31
  • 网络出版日期:  2023-02-10
  • 刊出日期:  2023-08-24

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