Testing of and Research into Drill Rod Joints for Long Hole Drilling
TESTING OF AND RESEARCH INTO DRILL ROD JOINTS
FOR LONG HOLE DRILLING
ACARP PROJECT 4039
Ian Gray
Bill Daniel
June 2000
ABSTRACT
This report contains information on the requirements for drill rods being used in long hole horizontal drilling in coal seams. It includes experimental and analytical studies of drill rod joints currently in use. It concludes with recommendations on which rods should be used and what modifications to design and drill rod joint use would enhance their strength and life.
As a result of the project a drill rod test facility has been established and sufficient knowledge has been gained to design drill rod threads to suit multiple applications.
Ian Gray
Bill Daniel
June 2000
INDEX
| Page No. | |
| ABSTRACT | 2 |
| INDEX | 3 |
| 1. OBJECTIVES | 4 |
| 2. BACKGROUND | 4 |
| 3. WIRELINE DRILL RODS | 4 |
| 4. REQUIREMENTS OF DRILL ROD JOINTS FOR LONG HOLE DRILLING | 6 |
| 5. PHYSICAL TESTING | 9 |
| 6. ANALYSIS | 10 |
| 6.1 Finite Element Studies | 11 |
| 6.2 Experimental Information | 12 |
| 6.3 Comparison with Experimental Measurements | 14 |
| 7. ALTERNATIVE TOOL JOINTS | 15 |
| 7.1 Thicker Tool Joints | 15 |
| 7.2 Better Material | 16 |
| 7.3 A More Clever Design | 17 |
| 8. WHICH ROD TO USE | 18 |
| 9. THE USE OF DRILL RODS | 19 |
| 10. CONCLUSIONS | 19 |
| 11. ACKNOWLEDGMENTS | 20 |
| 12. REFERENCES | 20 |
| 13. FIGURES | 21 |
1. OBJECTIVES
The objectives of this project were:
To develop a research facility to be able to test, analyse and and design drill rod joints.
To complete tests and analysis on the rods used in in-seam drilling, namely NT, NQ and CHD76.
To examine failed drill rods metallurgically.
The purpose of the project was to gain a better understanding of existing rod types and a knowledge of their behaviour so that better joints may be able to be designed to permit longer in-seam drilling.
2. BACKGROUND
ACARP Project C3023 "Optimisation of Long Hole Drilling Equipment", March 1994 (1), contained information on several in-seam boreholes from Tahmoor and Northcliff. These boreholes were analysed with the assistance of a torque and drag simulator and as a result good information was gained on drill rod loading. This report contained a number of recommendations which included the full testing of NT and NQ rod joints and that alternative drill rods should be located or be able to be designed. This project followed to answer the questions it contained.
3. WIRELINE DRILL RODS
The drill rods used for long hole drilling at the time of instigation of the project were all wireline drill rods. Two made by Longyear (now Boart Longyear), namely NQ and CHD76. A third rod was a Boyles NT rod.
Both the NQ and NT rods were manufactured out of 2.75" (69.85 mm) OD, 2.375" (60.32 mm) ID steel tube, probably from similar sources. The tube has a yield strength of approximately 90000 psi (621 MPa). Both rod types had the thread directly cut into the tube. In each case the thread is a near square thread which is cut at a taper into the tube. Hand tightening of the drill rod thread leads to the joint not closing fully. Some significant torque needs to be applied to the thread to make up the joint so that the box of the thread rides over the pin and impinges upon the buttress at the root of the pin. The NQ and NT rod joint details are shown in Figure 1 and 2 respectively. The NQ is a shorter thread of 3 tpi pitch and deeper thread form than the NT which has a pitch of 2 tpi.
The CHD76 rod has a rather different form. The rod joint is made of thicker material 0.295" (7.5 mm) than the rod body (0.188", 4.76 mm) but is uniform in outside diameter. The thread is also rather different having a slope on the leading edge of the thread. Essentially though the thread behaves in a similar manner as the NT and NQ designs. This thread detail is shown in Figure 3. The pitch is 2.5 tpi.
The behaviour of such tapered buttress threaded joints may be described as follows. The joint is screwed together and as it is made up contact is made between the inner and outer diameters of the box and pin respectively. Increased make up causes an interference fit between the pin and box as each rides over the other. Finally the end of the box intersects the root of the pin and a make up torque is applied which pre-loads one against the other.
The pre-load serves to help lock one against the other thus eliminating movement and fretting in the joints when under load.
Whilst these rods were principally designed for operation in tensile stress they behave well in compression as the buttress takes the compressive load substantially while the thread contacts remain together. The absence of the buttress would lead to the thread opening up and load transfer changing from one side of the threads to the other.
The interference fit on the taper helps ensure that the rod joint has the capacity to resist bending moments. The interference fit does however come at a price, namely that significant tensile and compressive hoop stresses may occur in the box and pin respectively. These hoop stresses are critically affected by the degree of interference between rod joints. As a result the tolerances achieved in manufacture are extremely important. Incorrect tolerance can easily lead to either yield on make-up or alternatively significantly reduced moment capacity of the joint. In the former case yielding of the material in the joint during make-up in one case may lead to too loose a fit when the next rod joint is made up. Repeated yielding leads to failure.
The effects of interference are worth discussion. Take for example an NQ rod with a 2 degree included angle taper on the thread. Assume a 3 mm standoff of the buttress at first make up. Then on completion of the thread the interference is 0.105 mm. Mid joint this interference is spread more or less evenly between the inner and outer sections resulting in a deformation of half the interference on each. This deformation corresponds to a hoop stress of 170 MPa and an interference stress of 12 MPa. At the thinner ends of the thread however this hoop stress increases significantly because the interference deformation principally occurs in the thinner section (tip) of the thread.
It can also be noted that the interference friction opposing the make up of such a thread can be very large. Depending on the coefficient of friction between pin and box the make up torque may not be adequate to permit pre-loading of the buttress. The correct cleanliness and lubrication of the joint is therefore extremely important.
The manufacturers recommended make up torques were considered both in the analytical and experimental aspects of the project. These recommendations varied through the life of the project but the values actually used are shown in Table 1.
| Rod | Make up Torque |
| NQ | 1.2 kNm |
| NT | 1.6 kNm |
| CHD76 | 3.0 kNm |
Table 1: Drill Rod Make Up Torques
4. REQUIREMENTS OF DRILL ROD JOINTS FOR LONG HOLE DRILLING
Long directional in-seam boreholes are never straight. There is a need to follow the varying rolls and faults within a coal seam. To achieve this drilling is undertaken with a bottom hole assembly containing a downhole mud motor. This has a characteristic angular build rate. There is a tendency to drill first in one direction and then the other to keep the hole on target in the seam. Because (too much) emphasis is placed on staying in-seam the bottom hole assembly is given a substantial bend so that it can be made to change trajectory quite suddenly at build rates up to 0.6 degree/m. Multiple changes of tool face angle lead to a hole with multiple dog legs.
Holes with multiple dog legs lead to excessive friction between the drill rods and the borehole wall. Not only is the build up rate important but so is the number of changes in direction. If a constant rate of angular build were maintained then the drill rods would only need be subject to a force couple at each end to generate a moment that would cause the drill rod to follow the curve. This can be readily demonstrated by flexing a ruler whilst holding it at each end and applying a couple between finger and thumb. A constant curve is developed easily. If however multiple flexures are generated in the ruler then more opposing force couples are required. The same applies to a drill rod.
A major factor limiting the length of near horizontal boreholes is buckling. When drill rods are pushed hard enough in a horizontal hole they tend to climb up the wall in a process that leads to helical buckling. Once helical buckling occurs the friction between drill rod and borehole wall rises very rapidly effectively limiting further drilling.
The load at which helical buckling occurs is a function of the drill rod stiffness, weight and the diameter of the borehole. Table 2 reproduced from the report on ACARP project C3023 describes the various buckling loads for NT/NQ and HT/HQ drill rods. The formulae from which these values are derived are also described therein.
Rod Borehole Helical Distance (m)
Type Diameter Buckling Coefficient of Coefficient of
(OD/ID) mm (in) Load kN fric - 0.17 fric - 0.25
mm/in
NT/NQ 88.9 77 kN 6902 4693
69.85/ (3.50)
60.32
2.75/ 95.3 66 kN 5977 4065
2.375 (3.75)
HT/HQ 101.2 180 kN 10819 7357
88.9/ (4.00)
77.8 108.0 147 kN 8832 6007
3.5/ (4.25)
3.06 114.3 127 kN 7650 5202
(4.50)
Table 2. Lengths to which drill rods can be slid into straight, horizontal, water filled boreholes before the onset of helical buckling.
Thus NQ/NT type drill rods require 77 kN thrust in a 88.9 mm diameter hole before the onset of buckling. This is the smallest hole that the rods are likely to be used in and correspondingly has the highest buckling load. Several other factors also need be taken into account. The first is the axial thrust capacity of the drill rig. The Longyear LMC75 drill can apply 135 kN of thrust. For the reasons described in the next paragraph this thrust should be limited.
The rods should be able to withstand more pullout load than that available to push them into the hole. The reason for this is that in a hole containing multiple dog legs the drill rods tend to lie on the insides of the bend. When axial load is applied the rods buckle slightly (not in the manner of helical buckling) away from the inside of the bend. Thus normal forces and hence friction between the rod and the borehole wall are reduced and the rod slides in. When the string is pulled out the opposite effect occurs and the normal force and friction between the drill rod and borehole wall increases. This results in pull out loads that may be much higher than those required to push the rods in. Once again the drill rig pullout capacity should be considered. In the case of the Longyear LMC75 this is 135 kN, the same as the thrust.
The bending moment that the drill rods may be subject to in a section of uniform rod is directly proportional to the curvature of the the rod. Angular build rates of 0.6 degrees per metre are fairly common but rise occasionally to 1.2 degrees/m. These correspond to bending moments of 1.14 kN-m and 2.28 kN-m respectively. Where the rods are connected to stiffer members locally higher bending moments may be developed than indicated by hole curvature. Special transition subs with a gradual change of stiffness are needed between different components of varying stiffness to limit the development of high local stresses.
The operational torque capacity required of drill rods should not exceed the make up torque that the joint can readily withstand. This should be in turn set as the maximum torque available from the drill rig whilst being used with that rod. The torque required to rotate a 2 km long NQ/NT drill string in a straight water filled horizontal borehole with a coefficient of friction of 0.25 between rod and coal is 1.14 kN-m. The presence of dog legs in the hole rapidly increases the torque required.
Examining two figures from the report on project C3073 is enlightening. These show the required torque and thrust to drill 2 km using NQ or NT drill rods. These are reproduced in Figures 4 and 5 respectively. The case considered is with a build-up rate of the BHA of 0.3 degrees/m, a bit torque of 400 Nm, a coefficient of friction of 0.25 and a hole diameter of 88.9 mm. The figures contain information with varying tool face angle change (flip flop) intervals and varying thrusts.
Examining Figure 4 it is possible to see that the maximum bit load that can be accommodated is 2.0 kN else the collar thrust exceeds the value of 77 kN associated with helical buckling. The tool face angle change rate must be limited to 18 m to be able to withdraw the drill string (110 kN) within the capacity of a Longyear LMC75 drill rig (135 kN). Examining Figure 5 it is possible to see that the corresponding torque to rotate the string is about 2.8 kNm. In the event that the drill string becomes stuck and cannot be directly pulled back the torque required to turn the string while pulling back is 4 kN-m. If drilling requires a higher bit load than 2 kN then either the hole must be drilled more smoothly (less BUR or fewer tool face changes) or be shorter. In either case the torque can be expected to remain at about 2.8 kNm. In reality the value of 2.0 kN bit load is too light to ensure a reasonable penetration rate and the more realistic value is 10 kN. This reduces the depth that can be reached to about 1.4 km as thrust and torque requirements rise rapidly.
The pressure inside the drill rods is substantially governed by that required to rotate the downhole motor. Normally this is about 500 psi (3.5 MPa), but may rise to 1000 psi (7 MPa) at times. The pressure is then limited by a relief valve on the pump.
From these facts we may thus arrive at a drill rod joint working loadings as shown in Table 3.
Axial Thrust 77.0 kN
Axial Pullout 110.0 kN
Bending Moment 2.3 kN-m
Torque 2.8 kN-m
Internal Pressure 7.0 MPa
Table 3. Maximum required working loads for N size rods in directional long hole drilling.
In the event of a jammed drill string the more axial pullout capacity and torque capability the better so as to permit the operator to pull and twist it out of the hole.
It can be noted that the only drill rod which has a torque capacity near that desired is the CHD76.
5. PHYSICAL TESTING
Figure 6 shows the test rig was designed and built that would permit sections of drill rod with joints at their centres to be loaded simultaneously in axial compressive load, torsion and bending. The torsion was applied to the rod through glued and clamped couplings. The gluing involved cleaning and priming and then using Loctite 680.
The drill rods were strain gauged with rosettes located on the inside of the pin and outsides of the box threads. These positions are shown in Figures 7, 8 and 9 for the NQ, NT and CHD76 drill rods respectively. Each strain gauge was monitored by a strain module. These each contained a 16 bit precision analogue to digital converter, microprocessor and power supply. They were monitored and operated from a PC via an RS485 link.
Loads were adjusted by the use of hydraulic jacks monitored by pressure gauges. In the case of the bending and torsion loads the jacks used were modified standard units whilst the axial load jack was a special hollow unit so as to permit its use in tension should the need arise. The axial thrust was applied through a cone and plate with a large ball bearing between. The bearing indented the plate to some degree with the result that bending moments were interfered with to some degree. To overcome this the drill rod bodies were strain gauged to permit direct measurement of bending moment and axial load. This procedure simplified the data gathering considerably.
Prior to making up the joint the output of strain gauges on the drill rods were measured and these values were used as a zero base from which other measurements were taken. The software written for the project was used to scan all strain gauges when loads were set and thus provide information on strain and hence stress. The software calculated the Von Mises Ratio (see below) and gave the operator an immediate idea of how close to yield the material at any strain gauge rosette had come. An endeavour was made to keep the Von Mises Ratio less than 0.65 ( VMR = 1.0 corresponds to yield).
It was fairly rapidly found that stress induced in the drill rod joint by make-up torque was quite significant. However once made up torsional stresses below this make up level did not significantly affect the drill rod joint stress level. This was useful as it permitted the variable of torsion to be eliminated from the test programme.
6. ANALYSIS
All analysis of physical test results and of numerical modelling are presented in terms of Von Mises Stress and Von Mises Ratio.
Von Mises Stress is the equivalent stress as defined in equation (1).
VMS = Sqrt(S1^2-S1S2+S2^2) (1)
where S1 and S2 are the principal stresses and VMS is the Von Mises Stress.
The Von Mises Ratio is defined in equation (2). It is the ratio of the Von Mises Stress to Yield stress based on tensile testing.
VMR = VMS/Sy (2)
where Sy is the yield stress (assumed to be 620 MPa for the materials used in the drill rods tested).
6.1 Finite Element Studies
The NT, NQ and CHD rods were modelled using meshes representing a typical cross-section along the axis. This was justified as a previous study in an undergraduate thesis project (2) had shown that the pattern of stress remained similar at different circumferential positions in axial tension or compression, and also that bending gave a similar distribution of stress through an axial cross-section normal to the neutral plane, to that obtained from tension or compression. Compression-only "gap" elements were used to join the pin to the box. These appear as lines joining the pin to the box on the deformed meshes. Radial interference was represented by causing artificial thermal expansion of these gap elements. Make up was represented in a similar way. The outer buttress was assumed to be in contact, with a contact pressure sufficient to balance the make up torque by generating a reaction torque at the buttress, using a coefficient of friction of 0.3. Gap elements joining the pin and box axially were given an artificial thermal expansion sufficient to cause the appropriate contact pressure. Make up torques used were the same as those used in the testing of the rods (NQ 1.2kN m, NT 1.6 kNm and CHD 3.0kNm).
With "perfect" thread geometry, the first turn of thread in the pin or the box near the buttress in contact is loaded substantially, causing local bending at the base of the groove of the tooth formation. The NQ and NT rods both showed local yielding near the surface at C and D in Figure 10. The CHD rod showed more substantial yielding, of about half the cross-section at C, using 620 MPa yield stress.
The stressing due to axial loading comparable to the highest loads applied during the testing was investigated. The presence of the make up torque reduces the proportion of applied bending or axial loading transmitted through the threads, as a substantial fraction is transferred across the buttress in contact, causing extra yielding around the buttress. This is beneficial in resisting fatigue failure, as changes in stressing of the thread are minimized. An overload of axial tension, however, can cause separation at the buttress, exposing the reduced cross-section at the base of the thread to the full axial loading. Without make up, the thread stressing changes more due to variation of bending or axial loading. Axial compression can cause yielding at the root of the box thread as shown at E in figure 10. This is more pronounced without make up.
The effect of torsional loading was also checked using a three dimensional finite element model. As also observed experimentally, the stresses induced due to torsion in operation are relatively small provided they are below the make-up torque.
Contour plots of representative examples of the predicted stressing in the rods are shown in Figures 11, 12 and 13. The first of these shows the effect of full make up torque on the outer buttress of an NQ rod, causing local yielding in compression on the inside of the box due to local bending near the buttress. The second figure shows the effect on the NT rod of make up plus axial compression of 108.6kN, the highest load applied in the testing. Substantial yielding has occurred in the box near the outer buttress, as well as partial yielding near the inner buttress. The overlap of the exaggerated deflected shape at the inner buttress, implies that the nodes there should have come into contact, relieving these stresses somewhat. The third figure shows the effect of axial compression of 322kN plus make up being applied to the CHD rod, and then the axial compression removed. This history of loading largely cancels the tensile stresses due to make up in the pin, and leaves local yielding on the inside of the box near the outer buttress, as marked. This would be the case if contact at the inner buttress does not take place. From a manufacturing viewpoint it is very difficult to ensure both buttresses contact simultaneously.
6.2 Experimental Information
The experimental measurements were processed to determine values of the ratio VMR = (Von Mises equivalent stress)/(Yield stress) measuring how close the material at a gauge is to yielding. As the torsional loading did not lead to high stresses, VMR values were found for various bending moments with the gauges up, putting them in tension due to bending, gauges down (in compression due to bending), and for a range of axial compressive loads.
Multi-linear regression was established for VMR versus axial load and bending moment at each gauge position with rods in the up and down position. The linearity of these results indicated that behaviour at the gauge sites was mainly elastic. The regression equation takes the form shown in equation (3).
VMR = a P + b M + c (3)
where P is the axial load and M is the bending moment
a, b and c are constants with c representing the make up value of VMR
The equations were evaluated and the gauge position located that would have led to the highest VMR. The axial and bending loads that would have led to a VMR = 1 (yield) are shown in Table 4.
Rod Worst Maximum Maximum
Type Gauge Axial Bending
Load (kN) Load (kN-m)
NQ 3 Box 290 5.8
NT 1 Box 330 7.2
CHD76 5 Pin 550 14.0
Table 4. Maximum Axial and Bending Loads for VMR = 1 at the Worst Gauge Position on the three rod types.
It should be noted that the stress levels at the worst gauge positions are higher than the stress levels in the critical locations of the thread, and the above values should be derated for service values to about 60% of those figures shown.
As expected the NQ and NT rods behaved somewhat similarly with the NT showing a higher bending moment consistent with its longer joint length. Both of these showed the maximum stress near the end of the box. The CHD 76 showed rather different behaviour with the root of the pin showing maximum stress.
The make up stresses were quite significant and are shown in Table 5.
Gauge NQ NT CHD76
Pin 1 0.138 0.118 0.129
Pin 2 0.194 0.151 0.274
Pin 3 0.272 0.125 0.528
Pin 4 0.174 0.228 0.284
Pin 5 0.159 0.029 0.234
Pin 6 0.071 0.064 -
Box 1 0.460 0.257 0.369
Box 2 0.276 0.223 0.270
Box 3 0.133 0.172 0.159
Box 4 0.077 0.048 0.069
Box 5 0.010 0.043 0.010
Box 6 0.015 0.028 -
Table 5. VMR's at make up for different gauge positions on the different rods.
The NT joint showed little change in VMR between tests. In addition the pure bending and pure axial tests showed consistent behaviour with the mixed load cases. This indicated there was little movement within the joint.
The NQ joint only showed consistent behaviour when loaded with a moment. Some slippage appeared to take place in the joint during the loading cycles.
The CHD 76 thread showed a significant redistribution of stress through the testing process. For example gauge 1 box initially showed a make up VMR of 0.442 which reduced to 0.369 with additional loading.
In the joints tested the compressive axial load tends to consistently increase VMR in the box but may not reduce the VMR in the pin thread, where the pre-existing stress due to make up is less significant.
The bending moment causes a similar pattern of stress to a compressive axial load in the compression side of the joint. On the tension side (eg gauges up), the VMR is reduced in the box where axial compression due to make up is significant. VMR may however increase elsewhere in the box when a moment loading alone is applied.
6.3 Comparison with Experimental Measurements
The finite element results can be viewed as either due to an axial load, or alternatively, as due to a bending moment producing the same maximum axial stress. The correlation of these results with the tests can be studied by estimating the stress at each position of a strain gauge rosette. Table 6 shows this for the NQ rod under bending load, using Von Mises' Ratios to report the stresses. In the case chosen remarkably good correlation exists for the pin behaviour but this is not the case with the box. The CHD and NT rods did not correlate as well, both showing a more even distribution of stress along the thread experimentally, than that predicted by the computer model, presumably due to yielding causing redistribution of stress, and due to imperfect thread geometry.
Strain Finite Element Experimental
Gauge Von Mises Ratio Von Mises Ratio
Pin 1 0.255 0.254
Pin 2 0.198 0.252
Pin 3 0.232 0.228
Pin 4 0.500 0.485
Pin 5 0.273 0.264
Box 1 0.231 0.110
Box 2 0.201 0.066
Box 3 0.141 0.185
Box 4 0.250 0.434
Box 5 0.239 0.326
Table 6. Comparison of Experimental and Predicted Stresses in the NQ Rod Joint Under Bending
7. ALTERNATIVE TOOL JOINTS
It is possible to make drill rods with tool joints that are as strong as the rod body. Three basic techniques can be used to achieve this:
- To make the tool joint thicker
- To make the joint out of better material
- To make the tool joint more cleverly
On the basis of information from Reference (1) there is certainly a need to make the tool joints capable of transmitting a higher torque. A spline transmits torque but can carry no axial load. Conversely a fine thread carries axial load and under torque tends to build up exceptionally high axial loads which tend to strip the thread.
Practical considerations are also important. The minimum number of thread turns that can be used on a thread before make up needs to be about four so as to avoid accidental unscrewing. There is also a need for a buttress pre-load that can be achieved with the desired make up torque.
Tolerance of manufacture is extremely important. Consider the case of the NQ rod discussed in Section 3. Here the interference between pin and box is 0.105 mm (0.00413"). Manufacturing tolerance could quite reasonably be expected to be +/- 0.025 mm (0.001"). Where a pin is undersize and the box oversize this could reduce the interference by 50%. Alternatively a large pin and small box could raise the interference by the same amount.
Re-assembly of a box swelled in a previous make up to a pin that has been slightly squeezed will lead to a fit with inadequate interference.
Poor cleaning in the drilling situation can lead to similar problems where foreign material trapped between box and pin swells the joint. This behaviour can also occur with heavy thread greases, particularly those with substantial solid inclusions. Thread lubricants should therefore enable low friction, prevent galling and should have a low viscosity so they are rapidly displaced out of the joint.
7.1 Thicker Tool Joints
Making tool joints thicker places greater demands on tolerance as the thicker joint will generate higher interference stresses for a given interference fit than a thin joint. High interference loads may lead to a failure to make up properly or galling during make up if the interference stress becomes excessive.
If the question of fit tolerance can be overcome then the joint made with thicker material will be stronger than a thinner one. There is another advantage to a thicker joint and that is the ease of making up the thread. In operation the low taper extremely shallow NT thread has caused some problems with cross threading whereas the CHD 76 thread is generally easily started without risk of cross threading.
The use of thicker tool joints has a complication in that either the rod material must be thicker or the tool joint must be made of material different from the rod body. Using a thicker rod raises the weight of the rod and reduces the distance it can be pushed before buckling occurs.
Using a different tool joint material involves joining a tool joint to the rod body. As all rod materials are high tensile steels they do not take kindly to welding.
The rod material can be welded conventionally but always at the cost of material strength. Oilfield drill rod manufacturers have for some time used with considerable success friction welding to joint the tool joint to the rod. This simply involves revolving the tool joint within the rod with sufficient interference, torque and thrust that the joint surface melts and a weld is created.
7.2 Better Material
There is always a trade off in engineering materials between high strength and ductility. It is possible to raise the strength of drill rod material but at the expense of the amount of deformation it will take before failure.
One compromise available to drill rod manufacturers is to raise the strength of material locally at the tool joint. This may be accomplished by a number of processes such as induction hardening. Boart Longyear have recently introduced the NRQ HP thread in which the joint has been strengthened by a hardening process to a quoted minimum yield strength of 1034 MPa (150000 psi). The rod material is otherwise the same as a normal NQ rod with a minimum yield strength of 620 MPa (90000 psi).
Such strength increases are a bonus but once again place a real burden on the manufacturer to ensure that manufacturing tolerance is maintained because the tool joint will not yield as readily to re-distribute manufacturing errors.
A thicker tool joint made of a non hardened steel can be re- cut to take up wear. Unfortunately high strength tool joints are also less readily re-machined.
7.3 A More Clever Design
The taper connection of drill rods used in current drilling is a successful design that works well. Several features can however be worked on to improve it. The first is to ensure that tightening of the thread does not lead to bulging of the box. Examination of the finite element stress analysis in Figures 11 to 13 will show the tendency for the joint to bulge in compression locally. Such deformation can frequently be seen on over tightened joints and takes the form of a spiral bulge over the box.
The flank angle of the thread in the tool joint is critically important to the bulge behaviour. Manufacturers of oilfield production tubing such as Grant Prideco have for some time recognised this phenomenon and have attempted to overcome it by hooking the mating threads so that they tend to pull in and resist the bulge.
Boart Longyear have apparently used this technique in the new NRQ HP drill rods. We do not have a drawing of this thread form at the moment to examine. The manufacturers quoted information on the NRQ HP and NQ is presented in Table 7.
Within the confines of a tool joint that must be machined into the rod body there is very little room to design alternative joint profiles. If a separate tool joint is fitted then the options include parallel threads and interference tapers before and after the thread. Such a tool joint becomes long but wear can be easily taken up by simply re-cutting the buttress position to develop the desired stand off.
NRQHP NQ
Rated Maximum Depth (m) 3000 1500
Rate Maximum Drilling Pullback (kN) 230 147
Rated Maximum Drilling Torque (kNm) 2.4 0.76
Burst Pressure (Box Shoulder) (MPa) 51.0 31.3
Midbody Minimum Yield Strength (MPa) 621. 621.
Midbody Minimum Tensile Strength (MPa) 723. 723.
Joint Minimum Yield Strength (MPa) 1034. 621.
Joint Minimum Tensile Strength (MPa) 1241. 723.
Pure Torsion to Failure (kNm) 8.5 5.2
Table 7. Comparison of Manufacturers Data on NQ and NRQHP rods
It had been hoped that an Atlas Bradford oilfield production tube would be made available for testing. This did not eventuate.
8. WHICH ROD TO USE
The rod type most generally used in directional drilling in coal mining is the Boart Longyear NQ. This rod has been demonstrated to perform well in drill holes up to 700 m. Theoretically it is somewhat deficient beyond this distance, most notably in the torque capacity of the drill rod joint.
Boart Longyear have recently introduced the NRQHP drill rod with considerably better specifications of joint performance despite being made of the same material. This improvement has been achieved by hardening the joints and improving their shape. These drill rods still have to be proven in drilling applications before there can be any certainty about their long term performance. Doubts specifically exist about the ductility of the joint and on machining tolerance. Provided these doubts are overcome the rod should prove a significant improvement though the torque capacity is still considered to be somewhat deficient for distances over 1300 m.
The NT drill rod has on paper better characteristics than the NQ but has been shown to suffer in some instances from lack of manufacturing tolerance. In addition it has been shown to be fairly easily cross threaded.
The CHD76 drill rod has been shown to have all the desirable characteristics of a drill rod for long hole drilling. It is only marginally weaker than desirable in loading under torque. This deficiency could be overcome by improving the tool joint by hardening or by modifying the thread profile to show the same hooked shape as the NRQ HP. Unfortunately the rod is a special order now and the market has a niche to be filled.
The writers preference in drill rods is to use a somewhat thicker more ductile tool joint with the possibility of surface hardening such as might be afforded by nitriding. Such surface hardening can only be achieved in a thicker tool joint so that a ductile core may remain.
If true long hole drilling is planned in the range of 2 km horizontal reach then the N size drill rods which have been the subject of this study are not suitable. A drill rod string with a greater stiffness to weight ratio is required. This means a drill rod of larger diameter but still with a slim wall. The 88.9 (3.5") OD, 77.8 (3.06") ID drill rod material used in HQ drill rods is admirable for this purpose but needs a suitable tool joint. The design of such a tool joint is a comparatively straight forward exercise following the study involved in this project.
The drilling of boreholes of 2 km or greater distance requires the use of bottom hole assemblies with lesser build up rates. These would also benefit from having a build characteristic that would permit the drill string to be rotated and drill straight ahead. Straight drilling minimises friction and rotation helps clear cuttings.
9. THE USE OF DRILL RODS
Drill rods must be manufactured to a tolerance range. The tolerance range that can be achieved will lead to a variation in stress developed on make up. Local yielding is likely to occur as pin and box threads deform to suit each other. This process is inevitable. It becomes a problem if rods are continually screwed together in different orders so that repeated plastic deformation takes place.
The solution to this is to keep using the rods in the same order so that the same pin and box are always made up. To facilitate the initial yielding it is recommended that new rods are made up several times to the maximum torque which they are likely to require assisted by a good lubricant. The rods should then be numbered and always used in the same order.
Such a process presupposes that the tool joint is made of an adequately ductile material to deform.
10. CONCLUSIONS
The most important conclusion of this study is that the only important torque on a tool joint is that applied at make up. Operational torques should be kept below this level and when this is achieved the stresses within the tool joint associated with these torques are insignificant.
An understanding of the torque requirements of a drill string has been an important contribution. So has the definition of the other loading capacity of the drill string.
The finite element modelling has helped in the visualization of the stress distribution within the joint. It has also demonstrated the soundness of hand calculation methods in joint design.
Little knowledge has been added to the knowledge of joints in combined axial load and bending with the exception of the CHD 76 tool joint.
The proposed procedure for maintaining a sequence of drill rod loading has the potential to significantly improve drill rod life and to reduce unexpected failure.
A level of understanding of tool joint design has been achieved so that other thread forms could be designed. Most design can be accomplished satisfactorily by hand methods. Hooked tooth forms may require recourse to numerical techniques such as finite element modelling.
Metallurgical failures are not apparently an issue. Chemicals do not seem to contribute to stress corrosion cracking in normal drilling. The only failed drill rod supplied was a copper beryllium rod which had been so mechanically abused that no metallurgical investigation was warranted.
The loading on a directional drill rod is in many ways far less demanding than that in rotary drilling where fatigue is important.
The drill rod test rig remains available for work. It has currently been modified for rotary fatigue testing of drill rod joints.
11. ACKNOWLEDGMENTS
The authors of this report thank ACARP for its support in undertaking this study. John Hanes and Jon Sleeman are to be congratulated on their tenacity in ensuring that a report has been produced. Undergraduate students Ross Evans, Gunther Wild and Eric Chua are thanked for their experimental contribution. The technical staff in the Department of Mechanical Engineering at the University of Queensland also made an important contribution by assisting student projects. Rod Young and Jim Evans also contributed significantly in the design and development of the strain gauging systems.
Boart Longyear are thanked for supplying a CHD 76 drill rod for testing as are Universal Drill Rigs for the supply of the NT rod.
12. REFERENCES
Evans Ross, (1994). Finite Element Analysis of Threaded Connectors Used in Coal Drilling. BE (Mech) Undergraduate Honours Thesis. The University of Queensland. October 1994.
Gray Ian, (1994). Optimisation of Long Hole Drilling Equipment. AMIRA. March 1994.
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Technical Papers Testing of and Research into Drill Rod Joints for Long Hole Drilling