Detection

The Top Drive assembly is composed of a 2-HP permanent magnet DC motor along with the drill string and bottom hole assembly. Figure 4.3 describes a simple electrical model for a permanent magnet DC motor. The characteristic response of such motors can be described by:

V _in

R L

V ω

Figure 4.3: Mathematical model of an electric Motor

T = K_ti (4.6)

V = Keθ ^{= Ked}θ

dt (4.7)

Where T is the motor torque in Nm, V_inis the induced voltage through the coils, I is the current, θ is the rotational displacement of the motor shaft in Radians, Kt

is the torque constant in Nm/A and K_eis the voltage constant in V /Rad/s. When an input voltage is applied V_in to the coils, the voltage equation is affected by the drop in the potential difference across the armature due to the coils electrical resistance.

This relationship is given by:

Vin = Rai + La

di

dt +V (4.8)

where R_ais the armature resistance in OhmsΩ, Lais the armature inductance in Henry (H) and i is the electrical current in Amperes (A). The inductance resis-tance is usually neglected it accounts for a fraction of the armature flux and is not

used in the generation of torque. A gear motor drives a mechanical load coupled with electrical static and dynamic mechanical subsystems such as the armature, cir-cuits, gears and shaft. The primary loads are the moment of inertia (J) and friction damping coefficient (ξ) of the motor’s shaft, therefore, the varying torque is given by:

T = Jd²θ

dt² +ξ^dθ dt + TL

We can write the following relationship equations after Equation 4.2 applying Newton’s and with Kirchhoff’s laws :

K_ti = Js²θ⁺ξθ^s V (s)− Ktθ^{s = R}ai + L_ais

it can be seen that i can be expressed as:

i =V− Ktsθ R_a+ L_as allowing to reduce equatio 4.2 to:

Jθ^s2+ξθ^{s = KV− Kt}θ^s

R_a+ L_as (4.9)

from Eq. 4.2 the relationship between the applied voltage (V ) and angular position (θ) can be derived as :

Gm(s) = θ^(s)

V (s)= V (s)

s[(R_a+ L_as)(Js +ξ^{) + K}t²] (4.10)

recalling thatω is the angular velocity given by ^d_dt^θ, equation 4.2 yields:

G_m(s) =ω^(s)

V (s) = V (s)

[(R_a+ L_as)(Js +ξ^{) + K}t²] (4.11) In state-space form, the governing equations above can be expressed by choos-ing the rotational speed and electric current as the state variables. The permanent magnet motor used in the rig is coupled with a gear mechanism to decrease the internal rotational speed of the motor and increase the maximum torque capacity.

The relationship between the internal motor speed and the outer shaft is called a gear ratio (G_r). Neglecting the inertia and friction losses of the pulley and adding the relationship of from Equation 4.1 we derive:

T_m(s)

V (s) = K_t(Js + L_a)

G_r((ξ^{+ Js)(L}as + R_a) + K_t²) (4.12) which relates the torque generated by the drawworks motor and the supplied voltage

Recalling from equation 4.1 the force acting along the drawworks line (F_Dw is

Kt

Figure 4.4: Block diagram of a DC motor

given by :

F_Dw= (F_{T d}+∆Lw)· csc[α⁺β^{] sin[}α^{] (4.13)}

Thus, the amount of torque being applied at the motor’s shaft will be given by:

T_m(s) = F_Dw· r sin(θ2) (4.14)

where r is the motor’s shaft radius in meters andθ2is the angle at which force is being exerted tangentially to the motors shaft and assuming that the cable does not slip and an ideal pulley system is employed, this is, considering the friction factors and inertias negligible then the downward velocity will equal the tangential velocity of the drawworks motor shaftω(s), and the amount of torque generated by the motor should be proportional to the force acting along the drawworks tension line. The resulting governing equation of motion assuming low spring coefficients and an over-damped response can be given by a simple second-second order transfer function.

Shear Stress and Pipe Torsional Deformation

Adequate selection and analysis of the drill string is critical to prevent per-formance limiters or pipe failure. Excessive vibrations can cause bit dysfunctions such as whirl, borehole patterns or stick-slip. A buckling and frequency response analysis of the drilling tube used in this study is presented in this section. The drilling string was specifically selected to increase the magnitude of vibrations and probability of failure if an inadequate control algorithm is present. The drill string chosen to run the drilling experiments discussed in this document was a 0.375”x0.035”x36” tube made of aluminum T 6061. Figure 4.5 pictures the cross section view of the aluminum pipe. Where r₁ and r₂ are the internal and external radius of the pipe respectively and τmax is the maximum tangential torsional stress that the pipe can endure without failing.

y

r

x

r

y

x

Figure 4.5: Tube cross section view

The maximum shear stress can be calculated assuming only an axial compres-sive force and static conditions by:

τ^{max=V Qmax}

I_cb (4.15)

Which yields a maximum torque of around 157.23 in-lb or 13.08 ft-lb incurring in an angular torsional deflection of 0.497 degrees per in-lbs of torque. Thus, a maximum torsional bending of 78.14 degrees can occur before mechanical failure.

Another approach to take into consideration both compression and torsional forces acting on the tube is given by the widely accepted yielding Von Mises yield stress criterion which is based on the maximum distortion energy theory. It relates triaxial stress to an equivalent stress that is a theoretical value to relate a general-ized three-dimensional (3D) stress state and compare it with the yield strength of a uniaxial failure criterion. If the triaxial stress exceeds the yield strength, a yield failure will occur. Under this assumptions, the maximum torsional stress was found to be around 20 - 25% less than under single axial load.

The area Moment of Inertia is given by:

I_m= π

64·(OD⁴−ID⁴) = π

64·(0.375in⁴−0.305in⁴) = 5.459E⁻⁴in⁴J_m=m

2 ∗(RO²+ RI²) (4.16) The theoretical axial spring constant for the hollow tube is given by:

K_c= Area· E

L = 0.00384845in²· 10⁷psi

36in = 1069.01lb f /in. (4.17) Clearly, the theoretical stiffness of the tube will yield an over-damped system especially if an axial force is acting on it. The tube will fail at a lower compressive stress due to buckling. The drilling string first natural frequency can be derived as:

ωⁿ⁼

√

123.71N/m

0.059928Kg = 45.4362Rad/s (4.18)