- Email: [email protected]
Fluidic DTH hammer with backward-impact-damping design for hard rock drilling
Journal of Petroleum Science and Engineering 171 (2018) 1077–1083
Contents lists available at ScienceDirect
Journal of Petroleum Science and Engineering journal homepage: www.elsevier.com/locate/petrol
Fluidic DTH hammer with backward-impact-damping design for hard rock drilling
T
Jianming Penga,b, Dong Gea,b, Xinxin Zhanga,c,d,∗, Maosen Wanga,b, Dongyu Wuc,d a
Key Laboratory of Drilling and Exploitation Technology in Complex Conditions, Ministry of Land and Resources, Changchun, 130026, China College of Construction Engineering, Jilin University, Changchun, 130026, China c Key Laboratory of Ministry of Education on Metallogenic Prediction of Nonferrous Metals and Geological Environment Monitoring, Central South University, Changsha, 410083, China d School of Geosciences and Info-Physics, Central South University, Changsha, 410083, China b
A R T I C LE I N FO
A B S T R A C T
Keywords: Hard rock drilling Fluidic oscillator Fluidic DTH hammer Impact body
Hard rock drilling is increasingly important due to progress in both the use of geothermal heat as a new energy source and the development of new unconventional gas fields. The DTH hammer drilling method is generally considered one of the best approaches for drilling hard rocks. A fluidic DTH hammer with backward-impactdamping design for hard rock drilling was proposed in the study. Striking features of this DTH hammer tool include extended life expectancy and capability to withstand drilling fluids solids content. A sequence of numerical and experimental investigations were performed on the newly designed fluidic DTH hammer. It is observed that the backward-impact-damping design is effective in protecting the fluidic oscillator from damage. Although the backward-impact-damping design may cause a decrease in the single-impact energy, the increased impact frequency can offset the aforementioned disadvantage to a certain extent. The paper brings attention to several interesting aspects of the fluidic DTH hammer. The accumulated data provide useful information to develop this type of a DTH hammer and facilitate its popularization and application in hard rock drilling.
1. Introduction
caused by high proportion of solids and abrasives in the mud system (Bybee, 2002; Carlos et al., 2003). With respect to its specific application and to focus on solving the aforementioned technical problems, development of mud hammers has been going on for some time but is still far from being satisfactory (Tibbitts et al., 2002). Until recent years, this situation has begun to change as an innovative DTH hammer termed as fluidic hammer has come up with a breakthrough in design conception. A fluidic hammer is a type of DTH hammer driven by mud and has proven efficient for exploration core drilling, geothermal drilling, and oil and gas drilling (Jian and Shang, 2005; Yin et al., 1996). During the German Continental Deep Drilling Program (Kontinentales Tiefborh programm der Bundesrepublik Deutschland, KTB), the fluidic hammer was considered as a potential hammer drilling tool that can be applied to super-deep hole drilling (Behr and Raleigh, 1990). Several field tests demonstrated that the ROP improved by more than 50% on an average in hard rock drilling with the use of a fluidic DTH hammer when compared with that in the conventional rotary drilling method (Jian et al., 2002; Yin et al., 1996).
Down-the-hole (DTH) hammer drilling with mud as opposed to air as the energy carrier is a competitive method for the production of drillholes (Tuomas, 2004). This technique is recognized for its improved rate of penetration (ROP) in hard formations and reduction of nonproductive time (NPT) associated with high consumption of drilling bits, excessive trips, stuck pipes, and fluid losses due to fractures (Santos et al., 2000; Staysko et al., 2011; Vieira et al., 2011). The use of mud can eliminate many of the disadvantages associated with airdriven DTH hammers (Tuomas and Nordell, 2000). A main difference is the capability to drill deep holes even in a water rich environment. Low energy consumption is another benefit. Furthermore, there are improvements in the working environment since dust is eliminated, and the atmosphere is oil-free and without grease residues. The immense potential commercial benefit of a DTH mud hammer motivated several development efforts over the last decades (Pixton and Hall, 1995). However, to date, the usage of mud hammers is hindered due to the technical difficulties associated with erosion and wear
∗ Corresponding author. Key Laboratory of Ministry of Education on Metallogenic Prediction of Nonferrous Metals and Geological Environment Monitoring, Central South University, Changsha, 410083, China. E-mail address: [email protected] (X. Zhang).
https://doi.org/10.1016/j.petrol.2018.08.046 Received 5 June 2018; Received in revised form 12 August 2018; Accepted 15 August 2018 Available online 17 August 2018 0920-4105/ © 2018 Elsevier B.V. All rights reserved.
Journal of Petroleum Science and Engineering 171 (2018) 1077–1083
J. Peng et al.
A fluidic oscillator is the core part of a fluidic hammer. It is based on the popular Coanda (1936) effect and characterized by the unsteady nature of an oscillating jet. The design of a fluidic oscillator is not completely new. Applications of such a device used in drilling engineering include an axial-oscillation tool used for reducing friction and torque in sliding drilling (He et al., 2015; Liu et al., 2017; Zhang et al., 2017), a submersible pump utilized in a pressure coring sampler (Zhang et al., 2016a), and a fluid-hammer tool for coiled-tubing (Castaneda et al., 2011; Livescu and Watkins, 2014; Livescu et al., 2017; Livescu and Craig, 2018). Given that a fluidic oscillator does not involve moving parts and possesses relative resistance to high temperatures, high pressure, and other harsh ambient conditions, a fluidic hammer can work more reliably in a drilled borehole when compared to tools with a moving or rotary mechanical valve (Carlos et al., 2003; Lehmann and Reich, 2013). Nevertheless, the utilization of the fluidic hammer is not so widespread as its obvious advantages would deserve. A few challenges still exist and should be addressed for the DTH hammer to achieve a distinct improvement and attract more industries to consider its use. The first challenge involves determining a material that is sufficiently strong to withstand high abrasive forces that result from drilling mud. Experiments with different materials indicate that tungsten carbide is the most suitable material for manufacturing a fluidic oscillator. However, in our previous design, an impact collision with considerable energy occurs when the piston reaches its rear stroke end position and poses a significant risk in terms of breaking the fluidic oscillator fabricated of brittle tungsten carbide material. In our first trial, the tungsten carbide fluidic oscillator only operated for approximately 46 s before it broke into pieces. Thus, the second challenge involved protecting the fluidic oscillator from destruction by the backward impact energy of the piston. Zhang et al. (2010) successfully applied cylindrical helical springs to buffer the backward impact of piston while adding wearing parts to the system and the results are less than satisfactory. In an effort to address the aforementioned problem, a fluidic DTH hammer with new design concept that delivers step-change in working performance and service life was proposed. The improved version of fluidic hammer tends to be a promising alternative to develop the new generation of mud hammers. In the study, the details of the design conception and a few relevant analyses are presented. 2. Design conception
Fig. 1. Schematic of the fluidic DTH hammer with backward-impact-damping design.
2.1. Operation principles and design concept
backward to its rear stroke end position, the impact collision between the piston and cylinder is undesirable since it can damage the fluidic oscillator composed of tungsten carbide. Based on the blocked load instability of the fluidic oscillator (referenced in a previous study), the supply jet can also be switched when the value of the motion velocity of the piston is close to zero (Zhang et al., 2018). Based on this jet switching mechanism, we introduced a specially designed structure in the vicinity of the rear stroke end position of the piston to produce a damping force that acts on the backward moving mass block to decelerate the motion of the piston to a stop or at least to a relatively low velocity without interrupting the supply jet to normally switch to the other side attachment wall. As shown in Fig. 1, the free stoke corresponds to the distance between the rear edge of the outermost cylindrical surface of the mass block and the front edge of the inner cylindrical surface of the bottom tie-in (which represents the section of the motion path that the mass block moves backwards on without or with little damping force). When the displacement of the piston exceeds the free stroke, most of the sectional area between the mass block and the inner surface of the bottom tie-in (which is for the passage of the fluid flow) is blocked and forms a blocked chamber with sudden increases in the pressure. The blocked chamber is analogous to the damping chamber of a hydraulic
As shown in Fig. 1, a fluidic DTH hammer consists of a fluidic oscillator that alternately directs fluid to the rear chamber of a cylinder and drives forward a piston combined with a mass block to strike the rear of the bit and then the front chamber of the cylinder through a port that bypasses the piston. The action allows the piston to move backwards to the rear stroke end and prepare itself for the subsequent forward stroke. While the piston moves, the supply jet in the fluidic oscillator is strongly attached to a side of the attachment walls. The load stability ensures the sustained and substantial entry of liquid into an operational chamber through the output channel such that it pushes the piston and makes it move. When the piston reaches the front end of its stroke and causes an impact on the bit, an intense pressure pulse charged with high pressure energy is formed in the rear chamber due to the sudden stop of the piston. A part of the pressure pulse propagates along the output channel to the feedback loop and produces a transverse force and flow momentum on the supply jet to trigger the supply jet to detach from the side wall to which it was initially attached. At this point, the oscillatory jet switch of the fluidic oscillator is attributed to the abrupt change in the flow field when a certain level of pressure is met in the output channels (Zhang et al., 2018). Conversely, when the main jet is switched and the piston moves 1078
Journal of Petroleum Science and Engineering 171 (2018) 1077–1083
J. Peng et al.
Fig. 2. Schematic of the backward stroke end damping regime. (a) Initial stage; (b) Second stage and (c) Third stage.
damper. The induced damping force acts on the integration of mass block and piston to decrease its velocity. With proper design of the backward-impact-damping structure, the velocity of the mass block in conjunction with the piston is significantly reduced and results in a lower level of backward impact energy of the piston or even a reverse movement of the piston prior to reaching its full stroke position. Predictably, the use of the damping structure to protect the fluidic oscillator is effective and especially when the jet switch occurs halfway during the backward movement of the piston. The resulting stroke of the piston is slightly shorter than the designed full stroke and completely avoids impact collision between the piston and cylinder.
Fig. 3. Force diagram of the impact body. The pressure at the outlet is equal to atmospheric pressure.
and is 9.81 m/s2, and Fdam denotes the damping force that occurs when the impact body is close to its rear stroke end position. Based on the working condition, the frictional force Ffp and buoyant force Fbuo are sufficiently low such that they can be ignored. Thus, the second and third items in the right side of the equation are omitted to simplify the calculation. Furthermore, the pressure-induced force that acts on the piston is calculated as follows:
2.2. Backward stroke end damping regime Fig. 2 shows a schematic of the backward stroke end damping regime of the mass block. The mass block moves freely with little damping force in the initial stage (Fig. 2 (a)). When the outermost cylindrical surface of the mass block is about to enter the inner central passage of the bottom tie-in, a sharp-edged orifice forms between the mass block and the bottom tie-in (Fig. 2 (b)). The liquid is removed through the sharp-edged orifice with high velocity. The mass block continues moving, and the sharp-edged orifice subsequently switches off, and an annular gap forms between the mass block and the bottom tie-in (Fig. 2 (c)). The liquid flow passing through the annular gap and the length of the annular gap increases with the backward movement of the mass block. During the process, given the compressibility of the liquid and the sudden contraction in the sectional area, high pressure emerges in the damping chamber and produces a damping force that acts on the mass block. The inner diameter of bottom tie-in Dt is constant; and thus, the effective action area of the mass block is adjusted by changing the value of the diameter of the mass block Dh.
Fpis = Pf Af − Pr Ar
(2)
Where Af and Ar denote the front and rear displacement areas, respectively, of the piston, and Pf and Pr denote the corresponding pressures. The annulus flow from the damping chamber to the outlet can be considered as a combined transient Couette–Poiseuille flow. The pressure difference is calculated by using the Darcy–Weisbach formula as follows (Baker, 1996):
ΔP = fd LU 2ρ /(2d )
(3)
Where ΔP denotes the pressure difference between the damping chamber and outlet. fd denotes the friction factor and are given by Eq. (4), assuming smooth turbulent flow. L denotes the annulus length, U denotes the mean flow velocity in the annular gap, d denotes the hydraulic diameter, and ρ denotes the fluid density. Specifically, d is calculated by dividing four times the cross-sectional area by the wetted perimeter.
2.3. Motion equations
fd = 0.316/Re0.25
Fig. 3 is a schematic diagram illustrating force balance conditions of the integration of mass block and piston (henceforth termed as the impact body). The force balance equation of the impact body is given by Newton's second law as follows:
where Re=Ud/υ is the Reynolds number andυis the kinematic viscosity.
mx¨ = Fpis + Fbuo − Ffp − mg − Fdam
Fdam = ΔPAd
(4)
The damping force acting on the mass block is then calculated as follows:
(1)
(5)
where Ad=Amas-Apr denotes the damping displacement area. Amas denotes the maximum cylindrical cross-sectional area of the mass block. Apr denotes the cross-sectional area of the piston rod.
Where m denotes the mass of the impact body and corresponds to 8 kg in the study, Fbuo denotes the buoyant force, x¨ denotes the acceleration of the impact body, Fpis denotes the pressure-induced forces acting on the piston, and Ffp denotes the force induced by the friction between the piston and the cylinder. Additionally, g denotes the gravity acceleration 1079
Journal of Petroleum Science and Engineering 171 (2018) 1077–1083
J. Peng et al.
Fig. 4. Schematic of the experimental tests.
3. Numerical and experimental approach Both numerical simulations and experimental measurements are performed for the fluidic DTH hammer with backward-impact-damping design to gain a better understanding of its internal flow behavior and perform a few analyses. The aforementioned understanding and analyses aid in the development of design methodologies for the next generation of fluidic DTH hammer and verify the feasibility of the new design concept. The numerical approach used in the study is based on the computational fluid dynamics (CFD). A dynamic mesh modeling technique and user-defined functions (UDFs) are used in the study. The UDFs are derived from the motion equations discussed in chapter two. Details of the numerical implementation of the method are specified in previous studies (He et al., 2015; Peng et al., 2013; Zhang et al., 2016b). The performance parameters of the fluidic DTH hammer in terms of impact velocity, impact frequency, and single-impact energy of the impact body are measured by a non-contact measuring method as described in the study by Zhang et al. (2016b) and are shown in Fig. 4. Additionally, a piston-rod dye method is adopted to obtain the actual stroke of the fluidic DTH hammer. As shown in the right portion of Fig. 4, the piston rod is dyed black, and a part of the back-dyed piston rod is scraped by the cylinder lid when the fluidic DTH hammer operates for a period of time on the bench. The length of the scraped area corresponds to the stroke of the fluidic DTH hammer.
Fig. 5. Velocity histories of the impact bodies for four cases with three different values of Rb, namely Rb = 0.964, Rb = 0.933, and Rb = 0.92 and the final case without damping. The results are obtained by numerical computations.
min, and the outer diameter of the Fluidic hammer is 86 mm. The diameters of the piston and piston rod are 25 and 18 mm, respectively. In the motion of the impact body without damping as delineated by the dash-dotted curve, the velocity increases with respect to time until the impact body reaches the full stroke end position and forms an impact at point a. Given the rebounding force, a reverse peak velocity occurs ephemerally, and the velocity then negatively increases with the forward movement of the impact body. As delineated by the other three curves when the blocking ratio varies, the impact bodies almost simultaneously enter the damping process (ignoring the marginal differences). The start damping moment is denoted as point d at the curves. This is potentially because their free strokes are equal. The marginal deviation is due to the difference of the sharp-edged-orifice stage when the displacement of impact bodies is close to the free stroke. As displayed in Fig. 5, d-b, d-c and d-e represent the damping process of three cases with different blocking ratio; the decrease rate of velocity in d-e section is higher than that of the others. Followed by the damping process, there exists a sudden change of velocity for the cases of Rb = 0.933 and Rb = 0.920. This is caused by the backward impacts of the impact bodies. The values of Vim1 and Vim2 denote their corresponding backward impact velocities; and they are all less than Vim0 (the impact velocity without damping). The results indicate that the backward impact velocity decreases with increases in the blocking ratio, and this is beneficial in preventing damage to the fluidic
4. Results and discussion The damping effect is mainly affected by the cross-sectional area of the annular gap between the mass block and the bottom tie-in. When the inner diameter of the bottom tie-in is constant, the cross-sectional area of the annular gap is determined by the diameter of the mass block. The cross-sectional area of the annular gap is calculated as follows:
Aann = Abt − Amas
(6)
Where Aann denotes the cross-sectional area of the annular gap, Abt denotes the cross-sectional area of the central passage of the bottom tiein. In order to non-dimensionalize the parameter and reflect a few blocking characteristics of the damping structure, an index of blocking ratio Rb is defined and expressed as follows:
Rb = 1 −
D2 Aann A = mas = h2 Abt Abt Dt
(7)
Fig. 5 shows the velocity histories of the impact bodies of four cases in approximately one impacting period. The supply flow rate is 200 L/ 1080
Journal of Petroleum Science and Engineering 171 (2018) 1077–1083
J. Peng et al.
Fig. 8. Variation in stroke Ls and single-impact energy E versus blocking ratio Rb. Fig. 6. Displacement histories of the impact bodies obtained by using the time–velocity integral of the data shown in Fig. 5.
constant value of blocking ratio, ζ decreases slightly when the Reynolds number increases, and ζ becomes almost constant when the blocking ratio is less than 0.933. The pressure loss gradually surpasses the flow inertia when the blocking ratio increases since the blocking ratio exceeds 0.933. Thus, the Euler number significantly increases with decreases in the Reynolds number. The increased damping effect is beneficial in decreasing the velocity of the impact body and stops its backward movement prior to reaching the full stroke end position. However, this also inevitably results in some loss of pressure in the impacting system. Fig. 8 shows the results obtain by CFD simulations and the experimental tests. It is observed that the numerical results obtained from the CFD simulations agree well with the experimental results, thereby exhibiting a satisfactory level of predictive accuracy. When the blocking ratio Rb exceeds 0.933, the stroke Ls decreases with increases in blocking ratio Rb, and the predicted data of the stroke are marginally lower than those in the experimental results. The aforementioned deviation is potentially because leakage from the cylinder lid was ignored in the CFD simulations. It can also be observed that the deviation exhibits an increasing trend when the blocking ratio Rb increases. This phenomenon may be attributed to the resultant increase of the leakage from the cylinder lid caused by the increased pressure difference in the damping chamber. As shown in Fig. 8, the variation in single-impact energy E is divided into two sections when the blocking ratio Rb increases. When the blocking ratio Rb is less than 0.933, the single-impact energy E varies within a negligible range. Nevertheless, when the blocking ratio Rb exceeds 0.933, the single-impact energy E decreases with increases in the blocking ratio Rb, and this corresponds to a significant trend. The variation trend in single-impact energy versus blocking ratio is similar to that of the stroke. It is concluded that shorter stroke results in lower single-impact energy, and this is consistent with the conclusions obtained in a previous study (Zhang et al., 2016b). As shown in Fig. 9, when the supply flow rate varies, the stroke is almost maintained as constant and is less than the full stroke 0.11 m while the impacting period increases with increases in the supply flow rate. It indicates that the variation in the supply flow rate does not significantly influence the measurement of stroke by piston-rod dye method. Additionally, as shown in Fig. 9, the forward impact velocity (negative maximum velocity) increases with increases in the supply flow rate. Therefore, the single-impact energy (which is proportional to the squared forward impact velocity) also exhibits an increasing tendency with increases in the supply flow rate. A general equation for the penetration rate of hammer drilling of rock is developed by Hustrulid and Fairhurst as follows (1971):
oscillator. When the value of Rb is equal to 0.946, the variation in motion velocity is relatively smooth as opposed to other cases in which a sudden change in velocity occurs. It is determined that the actual stroke in this case is shorter than the full stroke, and the impact body is hauled to a stop and moves in a reverse manner prior to reaching the full stroke end position. The point e as shown in Fig. 5 corresponds to the moment when the impact body is hauled to a stop by the damping force, and thus the value of Vt is 0. In this situation, the fluidic oscillator is maintained as intact without any damage caused by impact energy. The displacements curves shown in Fig. 6 are obtained by using time–velocity integral of the data presented in Fig. 5. The value of 0.11 m corresponds to the full stroke, and it indicates that the cases with Rb = 0.933 and Rb = 0.920, and the without damping case represents the situation in which the impact body can reach its full stroke end position in the backward movement regime. Conversely, the actual stroke of the case Rb = 0.949 is less than 0.11, and the variation in displacement around the maximum value is more gradual when compared with that in the other cases. Fig. 7 shows the relationship between pressure loss coefficient ζ (Euler number Eu) and Reynolds number Re where Re = 2U (Dt − Dh)/ υ . The pressure loss coefficient ζ is as follows:
ζ=
2ΔP ρU 2
(8)
As shown in the figure, the average magnitude of pressure loss coefficients ζ increases when the blocking ratio increases while the corresponding average magnitude of Reynolds numbers decrease. In a
Fig. 7. Pressure loss coefficient (Euler numbers) versus Reynolds number; the data are obtained from CFD simulations.
PR = 1081
EfT AS
(9)
Journal of Petroleum Science and Engineering 171 (2018) 1077–1083
J. Peng et al.
1) The backward-impact-damping design effectively slows the backward movement of the impact body to a stop or at least to a relatively small moving velocity without interrupting the supply jet to normally switch to the other side attachment wall. 2) When the blocking ratio is not less than a certain value (0.946 for this specific design in the paper), the stroke of the impact body is shorter than the full stroke. In this situation, the fluidic oscillator is maintained as intact without any damage due to impact energy. 3) The stroke of the impact body is almost maintained as constant with a variation in the supply flow rate, thereby indicating significant independence between them. 4) The backward-impact-damping design may cause a decrease in the performance of the fluidic DTH hammer; particularly the singleimpact energy is reduced with increases in the blocking ratio. However, this decrease in the performance can be limited to an acceptable level with proper design.
Fig. 9. Velocity and displacement histories of the impact bodies obtained by CFD simulations. The supply flow rate ranges from 180 L/min to 200 L/min; and the blocking ratio is 0.946.
Acknowledgment The study is supported by grant 201311112 from the Ministry of Land and Resources of China, research grant 2017M622609 from the Project funded by China Postdoctoral Science Foundation, and grant SXGJSF2017-5-(4-6) from the New Energy Program of Jilin Province, China. The authors sincerely acknowledge them. References Baker, R.C., 1996. An Introductory Guide to Industrial Flow. Mechanical Engineering Publications Limited, London ISBN 0852989830. Behr, H.J., Raleigh, C.B., 1990. Exploration of the deep continental crust. In: Deulsch, U., Marx, C., Rischmuller, H. (Eds.), Evaluation of Hammer Drill– Potential for KTB. Springer Verlag, Heidelberg, Berlin, pp. 310–321. Bybee, K., 2002. Mud-hammer drilling performance. J. Petrol. Technol. 54 (12), 38–87. https://doi.org/10.2118/1202-0038-JPT. Carlos, R.J., Carlos, V.M.A., Carvalho, D., Ronnie, H., Wajid, R., 2003. A new type of hydraulic hammer compatible with conventional drilling fluids. In: Presented at SPE Annual Technical Conference and Exhibition, Denver, Colorado, 5–8 October. SPE84355-MS. https://doi.org/10.2118/84355-MS. Castaneda, J., Schneider, C., Brunskill, D., 2011. Coiled tubing milling operations: successful application of an innovative variable water hammer extended-reach BHA to improve end load efficiencies of a PDM in horizontal wells. In: Presented at SPE/ ICoTA Coiled Tubing & Well Intervention Conference and Exhibition, the Woodlands, Texas, USA, 5-6 April. SPE-143346-MS. https://doi.org/10.2118/143346-MS. Coanda, H., 1936. Device for Deflecting a Stream of Elastic Fluid Projected into an Elastic Fluid. U.S. Patent 2,052,869, September 1. He, J.F., Yin, K., Peng, J.M., Zhang, X.X., Liu, H., Gan, X., 2015. Design and feasibility analysis of a fluidic jet oscillator with application to horizontal directional well drilling. J. Nat. Gas Sci. Eng. 27, 1723–1731. https://doi.org/10.1016/j.jngse.2015. 10.040. Hustrulid, W.A., Fairhurst, C., 1971. A theoretical and experimental study of the percussive drilling of rock. Part I theory of percussive drilling. Int. J. Rock Mech. Min. Sci. 8, 11–33. https://doi.org/10.1016/0148-9062(71)90045-3. Jian, Z., Shang, J., 2005. A type of advanced mud-hammer applied to oil drilling. In: Presented at SPE Asia Pacific Oil and Gas Conference and Exhibition, Jakarta, Indonesia, 5–7 April. SPE-93213-MS. https://doi.org/10.2118/93213-MS. Jian, Z., Sang, L., Hu, G., Zhang, W., 2002. The drilling experiment in geothermal well of DGSC-203 type hydro-efflux hammer. Geol. Prospect. 38 (5), 92–93. Lehmann, F., Reich, M., 2013. Development of alternative drive concepts for downhole hammers in deep drilling operations - a feasibility study. Oil Gas Eur. Mag. 39 (3), 119–123. Liu, Y., Chen, P., Ma, T., Wang, X., 2017. An evaluation method for friction-reducing performance of hydraulic oscillator. J. Petrol. Sci. Eng. 157, 107–116. https://doi. org/10.1016/j.petrol.2017.07.018. Livescu, S., Craig, S., 2018. A critical review of the coiled tubing friction-reducing technologies in extended-reach wells. Part 2: vibratory tools and tractors. J. Petrol. Sci. Eng. 166, 44–54. https://doi.org/10.1016/j.petrol.2018.03.026. Livescu, S., Watkins, T.J., 2014. Water hammer modeling in extended reach wells. In: Presented at SPE/ICoTA Coiled Tubing and Well Intervention Conference and Exhibition, the Woodlands, Texas, USA, 25-26 March. SPE-168297-MS. https://doi. org/10.2118/168297-MS. Livescu, S., Craig, S.H., Aitken, B., 2017. Fluid-hammer effects on coiled tubing friction in extended-reach wells. SPE J SPE-179100-PA. https://doi.org/10.2118/179100-PA. Peng, J.M., Zhang, Q., Li, G.L., Chen, J.W., Gan, X., He, J.F., 2013. Effect of geometric parameters of the bistable fluidic amplifier in the liquid-jet hammer on its threshold flow velocity. Comput. Fluids 82, 38–49. https://doi.org/10.1016/j.compfluid.2013. 05.002. Pixton, D., Hall, D., 1995. A New-generation Mud Hammer Drilling Tool Annual Report.
Fig. 10. Variations in impact frequency f and output power P based on the blocking ratio Rb.
where PR denotes the penetration rate, E denotes the single-impact energy, f denotes the impact frequency, T denotes the energy transfer coefficient, A denotes the drill-hole area, and S denotes specific energy of the rock. Specifically, the item Ef is replaced by the output power P of the fluidic DTH hammer, and thus Eq. (9) is rewritten as:
PR =
PT AS
(10)
Fig. 10 shows the variations in impact frequency f and output power P based on the blocking ratio Rb. The results are obtained by using a non-contact measuring system. As shown in the figure, the variation in the output power is in consistent with the frequency although the single-impact energy decreases with increases in the blocking ratio as shown in Fig. 8. When the value of the blocking ratio is 0.946, the increase in output power is relatively distinct. This is due to increases in impact frequency. A disadvantage of the backward-impact-damping design is that the single-impact energy is reduced although the increased impact frequency offsets the disadvantage to a certain extent. The decrease in output power is limited to within 5% with proper design that is at an acceptable level. 5. Conclusion In the study, we performed extensive numerical and experimental investigations on a fluidic DTH hammer with backward-impactdamping design for hard rock drilling. The investigations demonstrate the proposed tool is an innovative DTH hammer tool that deserves being better known in drilling engineering fields. The following conclusions are obtained in the study: 1082
Journal of Petroleum Science and Engineering 171 (2018) 1077–1083
J. Peng et al.
East Oil and Gas Show and Conference, Manama, Bahrain, 25–28 September. SPE140312-MS. https://doi.org/10.2118/140312-MS. Yin, K., Jian, Z., Jiang, R., 1996. Research on large impact energy hydro-hammer and its application. Explor. Eng. 4, 8–10. Zhang, Y.G., Peng, J.M., Liu, H., 2010. Buffering the backward impact of piston of liquidjet hammer. J. Jilin Univ. Earth Sci. Edit. 40 (6), 1415–1418. Zhang, X., Peng, J., Sun, M., Gao, Q., Wu, D., 2016a. Development of applicable ice valves for ice-valve-based pressure corer employed in offshore pressure coring of gas hydrate-bearing sediments. Chem. Eng. Res. Des. 111, 117–126. https://doi.org/10. 1016/j.cherd.2016.05.001. Zhang, X., Peng, J., Ge, D., Bo, K., Yin, K., Wu, D., 2016b. Performance study of a fluidic hammer controlled by an output-fed bistable fluidic oscillator. Appl. Sci-Basel 6 (10), 305. https://doi.org/10.3390/app6100305. Zhang, X., Peng, J., Liu, H., Wu, D., 2017. Performance analysis of a fluidic axial oscillation tool for friction reduction with the absence of a throttling plate. Appl. Sci.Basel 7 (4), 360. https://doi.org/10.3390/app7040360. Zhang, X., Zhang, S., Peng, J., Wu, D., 2018. A fluidic oscillator with concave attachment walls and shorter splitter distance for fluidic DTH hammers. Sens. Actuators A-Phys. 270, 127–135. https://doi.org/10.1016/j.sna.2017.12.043.
DOE project DEFG03-95ER82042, Novateck Inc., Provo, UT, USA. Santos, H., Placido, J.C.R., Oliveira, J.E., Gamboa, L., 2000. Overcoming hard rock drilling challenges. In: Presented at IADC/SPE Drilling Conference, New Orleans, Louisiana, 23–25 February. SPE-59182-MS. https://doi.org/10.2118/59182-MS. Staysko, R., Francis, B., Cote, B., 2011. Fluid hammer drives down well costs. In: Presented at SPE/IADC Drilling Conference and Exhibition, Amsterdam, The Netherlands, 1–3 March. SPE-139926-MS. https://doi.org/10.2118/139926-MS. Tibbitts, G., Long, R., Miller, B., Arnis, J., Black, A., 2002. World's first benchmarking of drilling mud hammer performance at depth conditions. In: Presented at IADC/SPE Drilling Conference, Dallas, Texas, 26–28 February. SPE-74540-MS. https://doi.org/ 10.2118/74540-MS. Tuomas, G., 2004. Effective use of water in a system for water driven hammer drilling. Tunn. Undergr. Space Technol. 19 (1), 69–78. https://doi.org/10.1016/j.tust.2003. 08.0011. Tuomas, G., Nordell, B., 2000. Down-hole water driven hammer drilling for BTES applications. In: Presented at Proceedings Terrastock 2000, 8th International Conference on Thermal Energy Storage, Stuttgart, Germany, pp. 503–508 28 August–1 September. Vieira, P., Lagrandeur, C., Sheets, K., 2011. Hammer drilling technology-the proved solution to drill hard rock formations in the Middle East. In: Presented at SPE Middle
1083
Contents lists available at ScienceDirect
Journal of Petroleum Science and Engineering journal homepage: www.elsevier.com/locate/petrol
Fluidic DTH hammer with backward-impact-damping design for hard rock drilling
T
Jianming Penga,b, Dong Gea,b, Xinxin Zhanga,c,d,∗, Maosen Wanga,b, Dongyu Wuc,d a
Key Laboratory of Drilling and Exploitation Technology in Complex Conditions, Ministry of Land and Resources, Changchun, 130026, China College of Construction Engineering, Jilin University, Changchun, 130026, China c Key Laboratory of Ministry of Education on Metallogenic Prediction of Nonferrous Metals and Geological Environment Monitoring, Central South University, Changsha, 410083, China d School of Geosciences and Info-Physics, Central South University, Changsha, 410083, China b
A R T I C LE I N FO
A B S T R A C T
Keywords: Hard rock drilling Fluidic oscillator Fluidic DTH hammer Impact body
Hard rock drilling is increasingly important due to progress in both the use of geothermal heat as a new energy source and the development of new unconventional gas fields. The DTH hammer drilling method is generally considered one of the best approaches for drilling hard rocks. A fluidic DTH hammer with backward-impactdamping design for hard rock drilling was proposed in the study. Striking features of this DTH hammer tool include extended life expectancy and capability to withstand drilling fluids solids content. A sequence of numerical and experimental investigations were performed on the newly designed fluidic DTH hammer. It is observed that the backward-impact-damping design is effective in protecting the fluidic oscillator from damage. Although the backward-impact-damping design may cause a decrease in the single-impact energy, the increased impact frequency can offset the aforementioned disadvantage to a certain extent. The paper brings attention to several interesting aspects of the fluidic DTH hammer. The accumulated data provide useful information to develop this type of a DTH hammer and facilitate its popularization and application in hard rock drilling.
1. Introduction
caused by high proportion of solids and abrasives in the mud system (Bybee, 2002; Carlos et al., 2003). With respect to its specific application and to focus on solving the aforementioned technical problems, development of mud hammers has been going on for some time but is still far from being satisfactory (Tibbitts et al., 2002). Until recent years, this situation has begun to change as an innovative DTH hammer termed as fluidic hammer has come up with a breakthrough in design conception. A fluidic hammer is a type of DTH hammer driven by mud and has proven efficient for exploration core drilling, geothermal drilling, and oil and gas drilling (Jian and Shang, 2005; Yin et al., 1996). During the German Continental Deep Drilling Program (Kontinentales Tiefborh programm der Bundesrepublik Deutschland, KTB), the fluidic hammer was considered as a potential hammer drilling tool that can be applied to super-deep hole drilling (Behr and Raleigh, 1990). Several field tests demonstrated that the ROP improved by more than 50% on an average in hard rock drilling with the use of a fluidic DTH hammer when compared with that in the conventional rotary drilling method (Jian et al., 2002; Yin et al., 1996).
Down-the-hole (DTH) hammer drilling with mud as opposed to air as the energy carrier is a competitive method for the production of drillholes (Tuomas, 2004). This technique is recognized for its improved rate of penetration (ROP) in hard formations and reduction of nonproductive time (NPT) associated with high consumption of drilling bits, excessive trips, stuck pipes, and fluid losses due to fractures (Santos et al., 2000; Staysko et al., 2011; Vieira et al., 2011). The use of mud can eliminate many of the disadvantages associated with airdriven DTH hammers (Tuomas and Nordell, 2000). A main difference is the capability to drill deep holes even in a water rich environment. Low energy consumption is another benefit. Furthermore, there are improvements in the working environment since dust is eliminated, and the atmosphere is oil-free and without grease residues. The immense potential commercial benefit of a DTH mud hammer motivated several development efforts over the last decades (Pixton and Hall, 1995). However, to date, the usage of mud hammers is hindered due to the technical difficulties associated with erosion and wear
∗ Corresponding author. Key Laboratory of Ministry of Education on Metallogenic Prediction of Nonferrous Metals and Geological Environment Monitoring, Central South University, Changsha, 410083, China. E-mail address: [email protected] (X. Zhang).
https://doi.org/10.1016/j.petrol.2018.08.046 Received 5 June 2018; Received in revised form 12 August 2018; Accepted 15 August 2018 Available online 17 August 2018 0920-4105/ © 2018 Elsevier B.V. All rights reserved.
Journal of Petroleum Science and Engineering 171 (2018) 1077–1083
J. Peng et al.
A fluidic oscillator is the core part of a fluidic hammer. It is based on the popular Coanda (1936) effect and characterized by the unsteady nature of an oscillating jet. The design of a fluidic oscillator is not completely new. Applications of such a device used in drilling engineering include an axial-oscillation tool used for reducing friction and torque in sliding drilling (He et al., 2015; Liu et al., 2017; Zhang et al., 2017), a submersible pump utilized in a pressure coring sampler (Zhang et al., 2016a), and a fluid-hammer tool for coiled-tubing (Castaneda et al., 2011; Livescu and Watkins, 2014; Livescu et al., 2017; Livescu and Craig, 2018). Given that a fluidic oscillator does not involve moving parts and possesses relative resistance to high temperatures, high pressure, and other harsh ambient conditions, a fluidic hammer can work more reliably in a drilled borehole when compared to tools with a moving or rotary mechanical valve (Carlos et al., 2003; Lehmann and Reich, 2013). Nevertheless, the utilization of the fluidic hammer is not so widespread as its obvious advantages would deserve. A few challenges still exist and should be addressed for the DTH hammer to achieve a distinct improvement and attract more industries to consider its use. The first challenge involves determining a material that is sufficiently strong to withstand high abrasive forces that result from drilling mud. Experiments with different materials indicate that tungsten carbide is the most suitable material for manufacturing a fluidic oscillator. However, in our previous design, an impact collision with considerable energy occurs when the piston reaches its rear stroke end position and poses a significant risk in terms of breaking the fluidic oscillator fabricated of brittle tungsten carbide material. In our first trial, the tungsten carbide fluidic oscillator only operated for approximately 46 s before it broke into pieces. Thus, the second challenge involved protecting the fluidic oscillator from destruction by the backward impact energy of the piston. Zhang et al. (2010) successfully applied cylindrical helical springs to buffer the backward impact of piston while adding wearing parts to the system and the results are less than satisfactory. In an effort to address the aforementioned problem, a fluidic DTH hammer with new design concept that delivers step-change in working performance and service life was proposed. The improved version of fluidic hammer tends to be a promising alternative to develop the new generation of mud hammers. In the study, the details of the design conception and a few relevant analyses are presented. 2. Design conception
Fig. 1. Schematic of the fluidic DTH hammer with backward-impact-damping design.
2.1. Operation principles and design concept
backward to its rear stroke end position, the impact collision between the piston and cylinder is undesirable since it can damage the fluidic oscillator composed of tungsten carbide. Based on the blocked load instability of the fluidic oscillator (referenced in a previous study), the supply jet can also be switched when the value of the motion velocity of the piston is close to zero (Zhang et al., 2018). Based on this jet switching mechanism, we introduced a specially designed structure in the vicinity of the rear stroke end position of the piston to produce a damping force that acts on the backward moving mass block to decelerate the motion of the piston to a stop or at least to a relatively low velocity without interrupting the supply jet to normally switch to the other side attachment wall. As shown in Fig. 1, the free stoke corresponds to the distance between the rear edge of the outermost cylindrical surface of the mass block and the front edge of the inner cylindrical surface of the bottom tie-in (which represents the section of the motion path that the mass block moves backwards on without or with little damping force). When the displacement of the piston exceeds the free stroke, most of the sectional area between the mass block and the inner surface of the bottom tie-in (which is for the passage of the fluid flow) is blocked and forms a blocked chamber with sudden increases in the pressure. The blocked chamber is analogous to the damping chamber of a hydraulic
As shown in Fig. 1, a fluidic DTH hammer consists of a fluidic oscillator that alternately directs fluid to the rear chamber of a cylinder and drives forward a piston combined with a mass block to strike the rear of the bit and then the front chamber of the cylinder through a port that bypasses the piston. The action allows the piston to move backwards to the rear stroke end and prepare itself for the subsequent forward stroke. While the piston moves, the supply jet in the fluidic oscillator is strongly attached to a side of the attachment walls. The load stability ensures the sustained and substantial entry of liquid into an operational chamber through the output channel such that it pushes the piston and makes it move. When the piston reaches the front end of its stroke and causes an impact on the bit, an intense pressure pulse charged with high pressure energy is formed in the rear chamber due to the sudden stop of the piston. A part of the pressure pulse propagates along the output channel to the feedback loop and produces a transverse force and flow momentum on the supply jet to trigger the supply jet to detach from the side wall to which it was initially attached. At this point, the oscillatory jet switch of the fluidic oscillator is attributed to the abrupt change in the flow field when a certain level of pressure is met in the output channels (Zhang et al., 2018). Conversely, when the main jet is switched and the piston moves 1078
Journal of Petroleum Science and Engineering 171 (2018) 1077–1083
J. Peng et al.
Fig. 2. Schematic of the backward stroke end damping regime. (a) Initial stage; (b) Second stage and (c) Third stage.
damper. The induced damping force acts on the integration of mass block and piston to decrease its velocity. With proper design of the backward-impact-damping structure, the velocity of the mass block in conjunction with the piston is significantly reduced and results in a lower level of backward impact energy of the piston or even a reverse movement of the piston prior to reaching its full stroke position. Predictably, the use of the damping structure to protect the fluidic oscillator is effective and especially when the jet switch occurs halfway during the backward movement of the piston. The resulting stroke of the piston is slightly shorter than the designed full stroke and completely avoids impact collision between the piston and cylinder.
Fig. 3. Force diagram of the impact body. The pressure at the outlet is equal to atmospheric pressure.
and is 9.81 m/s2, and Fdam denotes the damping force that occurs when the impact body is close to its rear stroke end position. Based on the working condition, the frictional force Ffp and buoyant force Fbuo are sufficiently low such that they can be ignored. Thus, the second and third items in the right side of the equation are omitted to simplify the calculation. Furthermore, the pressure-induced force that acts on the piston is calculated as follows:
2.2. Backward stroke end damping regime Fig. 2 shows a schematic of the backward stroke end damping regime of the mass block. The mass block moves freely with little damping force in the initial stage (Fig. 2 (a)). When the outermost cylindrical surface of the mass block is about to enter the inner central passage of the bottom tie-in, a sharp-edged orifice forms between the mass block and the bottom tie-in (Fig. 2 (b)). The liquid is removed through the sharp-edged orifice with high velocity. The mass block continues moving, and the sharp-edged orifice subsequently switches off, and an annular gap forms between the mass block and the bottom tie-in (Fig. 2 (c)). The liquid flow passing through the annular gap and the length of the annular gap increases with the backward movement of the mass block. During the process, given the compressibility of the liquid and the sudden contraction in the sectional area, high pressure emerges in the damping chamber and produces a damping force that acts on the mass block. The inner diameter of bottom tie-in Dt is constant; and thus, the effective action area of the mass block is adjusted by changing the value of the diameter of the mass block Dh.
Fpis = Pf Af − Pr Ar
(2)
Where Af and Ar denote the front and rear displacement areas, respectively, of the piston, and Pf and Pr denote the corresponding pressures. The annulus flow from the damping chamber to the outlet can be considered as a combined transient Couette–Poiseuille flow. The pressure difference is calculated by using the Darcy–Weisbach formula as follows (Baker, 1996):
ΔP = fd LU 2ρ /(2d )
(3)
Where ΔP denotes the pressure difference between the damping chamber and outlet. fd denotes the friction factor and are given by Eq. (4), assuming smooth turbulent flow. L denotes the annulus length, U denotes the mean flow velocity in the annular gap, d denotes the hydraulic diameter, and ρ denotes the fluid density. Specifically, d is calculated by dividing four times the cross-sectional area by the wetted perimeter.
2.3. Motion equations
fd = 0.316/Re0.25
Fig. 3 is a schematic diagram illustrating force balance conditions of the integration of mass block and piston (henceforth termed as the impact body). The force balance equation of the impact body is given by Newton's second law as follows:
where Re=Ud/υ is the Reynolds number andυis the kinematic viscosity.
mx¨ = Fpis + Fbuo − Ffp − mg − Fdam
Fdam = ΔPAd
(4)
The damping force acting on the mass block is then calculated as follows:
(1)
(5)
where Ad=Amas-Apr denotes the damping displacement area. Amas denotes the maximum cylindrical cross-sectional area of the mass block. Apr denotes the cross-sectional area of the piston rod.
Where m denotes the mass of the impact body and corresponds to 8 kg in the study, Fbuo denotes the buoyant force, x¨ denotes the acceleration of the impact body, Fpis denotes the pressure-induced forces acting on the piston, and Ffp denotes the force induced by the friction between the piston and the cylinder. Additionally, g denotes the gravity acceleration 1079
Journal of Petroleum Science and Engineering 171 (2018) 1077–1083
J. Peng et al.
Fig. 4. Schematic of the experimental tests.
3. Numerical and experimental approach Both numerical simulations and experimental measurements are performed for the fluidic DTH hammer with backward-impact-damping design to gain a better understanding of its internal flow behavior and perform a few analyses. The aforementioned understanding and analyses aid in the development of design methodologies for the next generation of fluidic DTH hammer and verify the feasibility of the new design concept. The numerical approach used in the study is based on the computational fluid dynamics (CFD). A dynamic mesh modeling technique and user-defined functions (UDFs) are used in the study. The UDFs are derived from the motion equations discussed in chapter two. Details of the numerical implementation of the method are specified in previous studies (He et al., 2015; Peng et al., 2013; Zhang et al., 2016b). The performance parameters of the fluidic DTH hammer in terms of impact velocity, impact frequency, and single-impact energy of the impact body are measured by a non-contact measuring method as described in the study by Zhang et al. (2016b) and are shown in Fig. 4. Additionally, a piston-rod dye method is adopted to obtain the actual stroke of the fluidic DTH hammer. As shown in the right portion of Fig. 4, the piston rod is dyed black, and a part of the back-dyed piston rod is scraped by the cylinder lid when the fluidic DTH hammer operates for a period of time on the bench. The length of the scraped area corresponds to the stroke of the fluidic DTH hammer.
Fig. 5. Velocity histories of the impact bodies for four cases with three different values of Rb, namely Rb = 0.964, Rb = 0.933, and Rb = 0.92 and the final case without damping. The results are obtained by numerical computations.
min, and the outer diameter of the Fluidic hammer is 86 mm. The diameters of the piston and piston rod are 25 and 18 mm, respectively. In the motion of the impact body without damping as delineated by the dash-dotted curve, the velocity increases with respect to time until the impact body reaches the full stroke end position and forms an impact at point a. Given the rebounding force, a reverse peak velocity occurs ephemerally, and the velocity then negatively increases with the forward movement of the impact body. As delineated by the other three curves when the blocking ratio varies, the impact bodies almost simultaneously enter the damping process (ignoring the marginal differences). The start damping moment is denoted as point d at the curves. This is potentially because their free strokes are equal. The marginal deviation is due to the difference of the sharp-edged-orifice stage when the displacement of impact bodies is close to the free stroke. As displayed in Fig. 5, d-b, d-c and d-e represent the damping process of three cases with different blocking ratio; the decrease rate of velocity in d-e section is higher than that of the others. Followed by the damping process, there exists a sudden change of velocity for the cases of Rb = 0.933 and Rb = 0.920. This is caused by the backward impacts of the impact bodies. The values of Vim1 and Vim2 denote their corresponding backward impact velocities; and they are all less than Vim0 (the impact velocity without damping). The results indicate that the backward impact velocity decreases with increases in the blocking ratio, and this is beneficial in preventing damage to the fluidic
4. Results and discussion The damping effect is mainly affected by the cross-sectional area of the annular gap between the mass block and the bottom tie-in. When the inner diameter of the bottom tie-in is constant, the cross-sectional area of the annular gap is determined by the diameter of the mass block. The cross-sectional area of the annular gap is calculated as follows:
Aann = Abt − Amas
(6)
Where Aann denotes the cross-sectional area of the annular gap, Abt denotes the cross-sectional area of the central passage of the bottom tiein. In order to non-dimensionalize the parameter and reflect a few blocking characteristics of the damping structure, an index of blocking ratio Rb is defined and expressed as follows:
Rb = 1 −
D2 Aann A = mas = h2 Abt Abt Dt
(7)
Fig. 5 shows the velocity histories of the impact bodies of four cases in approximately one impacting period. The supply flow rate is 200 L/ 1080
Journal of Petroleum Science and Engineering 171 (2018) 1077–1083
J. Peng et al.
Fig. 8. Variation in stroke Ls and single-impact energy E versus blocking ratio Rb. Fig. 6. Displacement histories of the impact bodies obtained by using the time–velocity integral of the data shown in Fig. 5.
constant value of blocking ratio, ζ decreases slightly when the Reynolds number increases, and ζ becomes almost constant when the blocking ratio is less than 0.933. The pressure loss gradually surpasses the flow inertia when the blocking ratio increases since the blocking ratio exceeds 0.933. Thus, the Euler number significantly increases with decreases in the Reynolds number. The increased damping effect is beneficial in decreasing the velocity of the impact body and stops its backward movement prior to reaching the full stroke end position. However, this also inevitably results in some loss of pressure in the impacting system. Fig. 8 shows the results obtain by CFD simulations and the experimental tests. It is observed that the numerical results obtained from the CFD simulations agree well with the experimental results, thereby exhibiting a satisfactory level of predictive accuracy. When the blocking ratio Rb exceeds 0.933, the stroke Ls decreases with increases in blocking ratio Rb, and the predicted data of the stroke are marginally lower than those in the experimental results. The aforementioned deviation is potentially because leakage from the cylinder lid was ignored in the CFD simulations. It can also be observed that the deviation exhibits an increasing trend when the blocking ratio Rb increases. This phenomenon may be attributed to the resultant increase of the leakage from the cylinder lid caused by the increased pressure difference in the damping chamber. As shown in Fig. 8, the variation in single-impact energy E is divided into two sections when the blocking ratio Rb increases. When the blocking ratio Rb is less than 0.933, the single-impact energy E varies within a negligible range. Nevertheless, when the blocking ratio Rb exceeds 0.933, the single-impact energy E decreases with increases in the blocking ratio Rb, and this corresponds to a significant trend. The variation trend in single-impact energy versus blocking ratio is similar to that of the stroke. It is concluded that shorter stroke results in lower single-impact energy, and this is consistent with the conclusions obtained in a previous study (Zhang et al., 2016b). As shown in Fig. 9, when the supply flow rate varies, the stroke is almost maintained as constant and is less than the full stroke 0.11 m while the impacting period increases with increases in the supply flow rate. It indicates that the variation in the supply flow rate does not significantly influence the measurement of stroke by piston-rod dye method. Additionally, as shown in Fig. 9, the forward impact velocity (negative maximum velocity) increases with increases in the supply flow rate. Therefore, the single-impact energy (which is proportional to the squared forward impact velocity) also exhibits an increasing tendency with increases in the supply flow rate. A general equation for the penetration rate of hammer drilling of rock is developed by Hustrulid and Fairhurst as follows (1971):
oscillator. When the value of Rb is equal to 0.946, the variation in motion velocity is relatively smooth as opposed to other cases in which a sudden change in velocity occurs. It is determined that the actual stroke in this case is shorter than the full stroke, and the impact body is hauled to a stop and moves in a reverse manner prior to reaching the full stroke end position. The point e as shown in Fig. 5 corresponds to the moment when the impact body is hauled to a stop by the damping force, and thus the value of Vt is 0. In this situation, the fluidic oscillator is maintained as intact without any damage caused by impact energy. The displacements curves shown in Fig. 6 are obtained by using time–velocity integral of the data presented in Fig. 5. The value of 0.11 m corresponds to the full stroke, and it indicates that the cases with Rb = 0.933 and Rb = 0.920, and the without damping case represents the situation in which the impact body can reach its full stroke end position in the backward movement regime. Conversely, the actual stroke of the case Rb = 0.949 is less than 0.11, and the variation in displacement around the maximum value is more gradual when compared with that in the other cases. Fig. 7 shows the relationship between pressure loss coefficient ζ (Euler number Eu) and Reynolds number Re where Re = 2U (Dt − Dh)/ υ . The pressure loss coefficient ζ is as follows:
ζ=
2ΔP ρU 2
(8)
As shown in the figure, the average magnitude of pressure loss coefficients ζ increases when the blocking ratio increases while the corresponding average magnitude of Reynolds numbers decrease. In a
Fig. 7. Pressure loss coefficient (Euler numbers) versus Reynolds number; the data are obtained from CFD simulations.
PR = 1081
EfT AS
(9)
Journal of Petroleum Science and Engineering 171 (2018) 1077–1083
J. Peng et al.
1) The backward-impact-damping design effectively slows the backward movement of the impact body to a stop or at least to a relatively small moving velocity without interrupting the supply jet to normally switch to the other side attachment wall. 2) When the blocking ratio is not less than a certain value (0.946 for this specific design in the paper), the stroke of the impact body is shorter than the full stroke. In this situation, the fluidic oscillator is maintained as intact without any damage due to impact energy. 3) The stroke of the impact body is almost maintained as constant with a variation in the supply flow rate, thereby indicating significant independence between them. 4) The backward-impact-damping design may cause a decrease in the performance of the fluidic DTH hammer; particularly the singleimpact energy is reduced with increases in the blocking ratio. However, this decrease in the performance can be limited to an acceptable level with proper design.
Fig. 9. Velocity and displacement histories of the impact bodies obtained by CFD simulations. The supply flow rate ranges from 180 L/min to 200 L/min; and the blocking ratio is 0.946.
Acknowledgment The study is supported by grant 201311112 from the Ministry of Land and Resources of China, research grant 2017M622609 from the Project funded by China Postdoctoral Science Foundation, and grant SXGJSF2017-5-(4-6) from the New Energy Program of Jilin Province, China. The authors sincerely acknowledge them. References Baker, R.C., 1996. An Introductory Guide to Industrial Flow. Mechanical Engineering Publications Limited, London ISBN 0852989830. Behr, H.J., Raleigh, C.B., 1990. Exploration of the deep continental crust. In: Deulsch, U., Marx, C., Rischmuller, H. (Eds.), Evaluation of Hammer Drill– Potential for KTB. Springer Verlag, Heidelberg, Berlin, pp. 310–321. Bybee, K., 2002. Mud-hammer drilling performance. J. Petrol. Technol. 54 (12), 38–87. https://doi.org/10.2118/1202-0038-JPT. Carlos, R.J., Carlos, V.M.A., Carvalho, D., Ronnie, H., Wajid, R., 2003. A new type of hydraulic hammer compatible with conventional drilling fluids. In: Presented at SPE Annual Technical Conference and Exhibition, Denver, Colorado, 5–8 October. SPE84355-MS. https://doi.org/10.2118/84355-MS. Castaneda, J., Schneider, C., Brunskill, D., 2011. Coiled tubing milling operations: successful application of an innovative variable water hammer extended-reach BHA to improve end load efficiencies of a PDM in horizontal wells. In: Presented at SPE/ ICoTA Coiled Tubing & Well Intervention Conference and Exhibition, the Woodlands, Texas, USA, 5-6 April. SPE-143346-MS. https://doi.org/10.2118/143346-MS. Coanda, H., 1936. Device for Deflecting a Stream of Elastic Fluid Projected into an Elastic Fluid. U.S. Patent 2,052,869, September 1. He, J.F., Yin, K., Peng, J.M., Zhang, X.X., Liu, H., Gan, X., 2015. Design and feasibility analysis of a fluidic jet oscillator with application to horizontal directional well drilling. J. Nat. Gas Sci. Eng. 27, 1723–1731. https://doi.org/10.1016/j.jngse.2015. 10.040. Hustrulid, W.A., Fairhurst, C., 1971. A theoretical and experimental study of the percussive drilling of rock. Part I theory of percussive drilling. Int. J. Rock Mech. Min. Sci. 8, 11–33. https://doi.org/10.1016/0148-9062(71)90045-3. Jian, Z., Shang, J., 2005. A type of advanced mud-hammer applied to oil drilling. In: Presented at SPE Asia Pacific Oil and Gas Conference and Exhibition, Jakarta, Indonesia, 5–7 April. SPE-93213-MS. https://doi.org/10.2118/93213-MS. Jian, Z., Sang, L., Hu, G., Zhang, W., 2002. The drilling experiment in geothermal well of DGSC-203 type hydro-efflux hammer. Geol. Prospect. 38 (5), 92–93. Lehmann, F., Reich, M., 2013. Development of alternative drive concepts for downhole hammers in deep drilling operations - a feasibility study. Oil Gas Eur. Mag. 39 (3), 119–123. Liu, Y., Chen, P., Ma, T., Wang, X., 2017. An evaluation method for friction-reducing performance of hydraulic oscillator. J. Petrol. Sci. Eng. 157, 107–116. https://doi. org/10.1016/j.petrol.2017.07.018. Livescu, S., Craig, S., 2018. A critical review of the coiled tubing friction-reducing technologies in extended-reach wells. Part 2: vibratory tools and tractors. J. Petrol. Sci. Eng. 166, 44–54. https://doi.org/10.1016/j.petrol.2018.03.026. Livescu, S., Watkins, T.J., 2014. Water hammer modeling in extended reach wells. In: Presented at SPE/ICoTA Coiled Tubing and Well Intervention Conference and Exhibition, the Woodlands, Texas, USA, 25-26 March. SPE-168297-MS. https://doi. org/10.2118/168297-MS. Livescu, S., Craig, S.H., Aitken, B., 2017. Fluid-hammer effects on coiled tubing friction in extended-reach wells. SPE J SPE-179100-PA. https://doi.org/10.2118/179100-PA. Peng, J.M., Zhang, Q., Li, G.L., Chen, J.W., Gan, X., He, J.F., 2013. Effect of geometric parameters of the bistable fluidic amplifier in the liquid-jet hammer on its threshold flow velocity. Comput. Fluids 82, 38–49. https://doi.org/10.1016/j.compfluid.2013. 05.002. Pixton, D., Hall, D., 1995. A New-generation Mud Hammer Drilling Tool Annual Report.
Fig. 10. Variations in impact frequency f and output power P based on the blocking ratio Rb.
where PR denotes the penetration rate, E denotes the single-impact energy, f denotes the impact frequency, T denotes the energy transfer coefficient, A denotes the drill-hole area, and S denotes specific energy of the rock. Specifically, the item Ef is replaced by the output power P of the fluidic DTH hammer, and thus Eq. (9) is rewritten as:
PR =
PT AS
(10)
Fig. 10 shows the variations in impact frequency f and output power P based on the blocking ratio Rb. The results are obtained by using a non-contact measuring system. As shown in the figure, the variation in the output power is in consistent with the frequency although the single-impact energy decreases with increases in the blocking ratio as shown in Fig. 8. When the value of the blocking ratio is 0.946, the increase in output power is relatively distinct. This is due to increases in impact frequency. A disadvantage of the backward-impact-damping design is that the single-impact energy is reduced although the increased impact frequency offsets the disadvantage to a certain extent. The decrease in output power is limited to within 5% with proper design that is at an acceptable level. 5. Conclusion In the study, we performed extensive numerical and experimental investigations on a fluidic DTH hammer with backward-impactdamping design for hard rock drilling. The investigations demonstrate the proposed tool is an innovative DTH hammer tool that deserves being better known in drilling engineering fields. The following conclusions are obtained in the study: 1082
Journal of Petroleum Science and Engineering 171 (2018) 1077–1083
J. Peng et al.
East Oil and Gas Show and Conference, Manama, Bahrain, 25–28 September. SPE140312-MS. https://doi.org/10.2118/140312-MS. Yin, K., Jian, Z., Jiang, R., 1996. Research on large impact energy hydro-hammer and its application. Explor. Eng. 4, 8–10. Zhang, Y.G., Peng, J.M., Liu, H., 2010. Buffering the backward impact of piston of liquidjet hammer. J. Jilin Univ. Earth Sci. Edit. 40 (6), 1415–1418. Zhang, X., Peng, J., Sun, M., Gao, Q., Wu, D., 2016a. Development of applicable ice valves for ice-valve-based pressure corer employed in offshore pressure coring of gas hydrate-bearing sediments. Chem. Eng. Res. Des. 111, 117–126. https://doi.org/10. 1016/j.cherd.2016.05.001. Zhang, X., Peng, J., Ge, D., Bo, K., Yin, K., Wu, D., 2016b. Performance study of a fluidic hammer controlled by an output-fed bistable fluidic oscillator. Appl. Sci-Basel 6 (10), 305. https://doi.org/10.3390/app6100305. Zhang, X., Peng, J., Liu, H., Wu, D., 2017. Performance analysis of a fluidic axial oscillation tool for friction reduction with the absence of a throttling plate. Appl. Sci.Basel 7 (4), 360. https://doi.org/10.3390/app7040360. Zhang, X., Zhang, S., Peng, J., Wu, D., 2018. A fluidic oscillator with concave attachment walls and shorter splitter distance for fluidic DTH hammers. Sens. Actuators A-Phys. 270, 127–135. https://doi.org/10.1016/j.sna.2017.12.043.
DOE project DEFG03-95ER82042, Novateck Inc., Provo, UT, USA. Santos, H., Placido, J.C.R., Oliveira, J.E., Gamboa, L., 2000. Overcoming hard rock drilling challenges. In: Presented at IADC/SPE Drilling Conference, New Orleans, Louisiana, 23–25 February. SPE-59182-MS. https://doi.org/10.2118/59182-MS. Staysko, R., Francis, B., Cote, B., 2011. Fluid hammer drives down well costs. In: Presented at SPE/IADC Drilling Conference and Exhibition, Amsterdam, The Netherlands, 1–3 March. SPE-139926-MS. https://doi.org/10.2118/139926-MS. Tibbitts, G., Long, R., Miller, B., Arnis, J., Black, A., 2002. World's first benchmarking of drilling mud hammer performance at depth conditions. In: Presented at IADC/SPE Drilling Conference, Dallas, Texas, 26–28 February. SPE-74540-MS. https://doi.org/ 10.2118/74540-MS. Tuomas, G., 2004. Effective use of water in a system for water driven hammer drilling. Tunn. Undergr. Space Technol. 19 (1), 69–78. https://doi.org/10.1016/j.tust.2003. 08.0011. Tuomas, G., Nordell, B., 2000. Down-hole water driven hammer drilling for BTES applications. In: Presented at Proceedings Terrastock 2000, 8th International Conference on Thermal Energy Storage, Stuttgart, Germany, pp. 503–508 28 August–1 September. Vieira, P., Lagrandeur, C., Sheets, K., 2011. Hammer drilling technology-the proved solution to drill hard rock formations in the Middle East. In: Presented at SPE Middle
1083