Development of new rotation concept for rock drills

Development of new rotation concept for rock drills

AKEPATI BHASKAR REDDY

Master of Science Thesis Stockholm, Sweden 2015

Development of new rotation concept for rock drills

Akepati Bhaskar Reddy

Master of Science Thesis MMK 2015:99 MKN 147 KTH Industrial Engineering and Management

Machine Design SE-100 44 STOCKHOLM

Examensarbete MMK 2015:99 MKN 147

Utveckling av nya rotationskoncept för bergborrar

Akepati Bhaskar Reddy

Godkänt

2015-09-08

Examinator

Ulf Sellgren

Handledare

Stefan Björklund

Uppdragsgivare

Atlas Copco Rock Drills AB

Kontaktperson

Anders Olson

Sammanfattning

Vid spränghålsborrning är det önskvärt att borra så nära tunnels periferi som möjligt. Det minsta avståndet från kanten definieras av borrmaskinens radie. Majoriteten av dagens borrmaskiner har en rotationsmotor som roterar en axel parallell till nackadaptern, vilket i sin tur roterar nackadaptern genom en drevuppsättning. Detta gör borrmaskinen otymplig.

Detta examensarbete, som utfördes på Atlas Copco Rock Drills AB, Örebro, undersöker rotationsmekanismen hos en borrmaskin. Projektets huvuduppgift var att utveckla alternativa koncept till rotationsmekanismen som använder nackadaptern som en del av rotationsmekanismen och reducerar borrmaskinens otymplighet.

För att hitta ett lämpligt alternativ genererades åtta olika koncept för en hydraulisk motor (med och utan transmission) som sedan utvärderades. Två koncept, 1) ”multi-kam vingmotor” och 2)

”hydrauliskt driven töjningsvåg-växel”, valdes för ytterligare funktionell konstruktion. Båda koncepten konstruerades under ideala förhållanden (100% effektivitet) för att uppnå det deplacement som krävs.

För ”multi-kam vingmotor” utfördes ett flertal konstruktionsförbättringar och effekten av olika konstruktionsparametrar analyserades. Olika varianter av motorn togs fram samt analyserades. En grov design genomfördes för ”hydrauliskt driven töjningsvåg-växel”, vilken valdes på grund av att designen inte förekommit I någon litteratur. CAD-modeller för båda koncepten samt relaterade varianter togs fram för att föreslå monteringslayouter och ventilmekanismer.

De två koncepten skulle reducera otympligheten hos borrmaskinen. För –och nackdelarna hos de olika varianterna har diskuterats. Koncepten måste utvecklas ytterligare för att kunna implementeras i en borrmaskin.

Nyckelord: hydraulisk motor, borrmaskin, vridning

Master of Science Thesis MMK 2015:99 MKN 147

Development of new rotation concept for rock drills

Akepati Bhaskar Reddy

Approved

2015-09-08

Examiner

Ulf Sellgren

Supervisor

Stefan Björklund

Commissioner

Atlas Copco Rock Drills AB

Contact person

Anders Olson

Abstract

In blast hole drilling, it is desirable to be able to drill as close as possible to the edge of the tunnel.

The minimum distance from the edge is defined by the radial size of the rockdrill. Most of the rock drills used today have a rotation motor that rotates an axel parallel to the shank, which further rotates the shank through a gear set. Thus making the rock drill bulky.

This thesis project carried out at Atlas Copco Rock Drills AB, Örebro, deals with the rotation mechanism of a rock drill. The main task of the project was to develop alternate concepts for rotation mechanism that would use the shank as a part of rotation mechanism and reduce the bulkiness of the rock drill.

In order to find a suitable alternative, eight different concepts for hydraulic motor (with or without transmission) were generated and evaluated against each other. Two concepts, 1) multi-cam vane motor concept and 2) strain wave hydraulic gear motor concept, were selected for further functional design. Both concepts were designed at ideal conditions (100% efficiency) to achieve the required displacement.

For the multi-cam vane motor, various design improvements were performed and the effect of different design parameters were also analyzed. Different variants of the motor were developed and analyzed. A rough design was performed for the strain wave hydraulic gear motor concept which was chosen for its novelty. CAD models for both the concepts and the related variants were developed for suggesting assembly layouts and valve mechanisms.

The two concept designs would reduce the bulkiness of the rock drill. The benefits and drawbacks of the different variants have been discussed. The concepts must be further developed for implementation into a rockdrill.

Keywords: hydraulic motor, rock drill, rotation

FOREWORD

In this section, the help, assistance, cooperation and inspiration from others has been acknowledged.

I would like to express my gratitude to my supervisors at Atlas Copco Rock Drills AB, Maria Pettersson and Erik Jakobsson, for the constant engagement, useful comments and remarks throughout this master thesis. Furthermore I would like to thank Anders Olson, R&D manager, Rock drills & rotation units, Rocktec division, for providing me an opportunity to work on this project.

This thesis would not have been possible without the help from my university supervisor, Stefan Björklund, and MSc project coordinator, Ulf Sellgren.

I would like to thank everyone at Atlas Copco Rock Drills AB and KTH Royal Institute of Technology, who have, directly or indirectly, helped me achieve my goals.

Finally, all thanks to almighty and my parents for their support and blessings, without which nothing would have been possible.

Akepati Bhaskar Reddy Stockholm, September 2015

NOMENCLATURE

In this section the symbols and abbreviations used in the report are listed. The symbols are listed in the order of their first appearance in the report.

Notations

Symbol Description Unit

nD Required rotation speed rpm

DB Drill bit diameter mm

f Impact frequency Hz

zB Bit movement per blow mm

Pin Input fluid pressure Pa

Pout Return fluid pressure Pa

Q Flow rate m³/min

 

^r, Polar coordinate system

RO Radius of mean circle of cam profile (sine curve) m

a Amplitude of sine curve m

n Twice the number of cams

k Number of vanes

B Width of motor unit m

W p Width of vane m

r p Head radius of vane m

 Angle of rotation rad



XC^,YC



Coordinates of point of contact



XR^,YR



Coordinates of centre of circular part of vane

 Angle of rotation of contact point rad

mn Slope of normal to the sine profile

 Angle formed by the line joining the centre of circular part of the vane to the point of contact on the sine profile, with the horizontal axis

rad

RC Distance of point of contact from the origin m

RR Distance of centre of circular part of the vane from the origin m

 Angle formed by line joining centre of circular part of vane and the point of contact, with the axis of vane

rad

 , Angles in the depiction of contact geometry (Figure 26 and Figure 63)

rad

RI Radius of circular profile (rotor or stator) m

1, 2

  Angular position of vane 1, 2 rad

chamber

A Area of chamber m²

, ,

a b c

A A A Parts of chamber area m²

1, 2, 3

b b b

A A A Parts of area component A_b m²

1, 2

R R

R R R_Rvalues for vane 1,2 corresponding to angular position  ₁, ₂ respectively

m

1, 2

c c

A A Parts of area component A_c m²

1_1, 1_ 2

c c

A A Parts of area component A_c1 m²

2 _1, 2 _ 2

c c

A A Parts of area component A_c2 m²

1, 2

  Angle of contact for vane 1,2 corresponding to angular position

1, 2

  rad

chamber

A Change in area of chamber m²

A_{  }d_

 Total change in area for small rotation d ^m2

dA Change in area for one sub-cycle of rotation m²

dV Volume of fluid added m³

i Instantaneous speed rpm

, , ,

a b c d

T T T T Torque constituents per unit width Nm/m

CX, CY

F F Contact force (per unit width) in direction normal to vane axis and parallel to vane axis respectively

N/m

 Half angle subtended by width of the vane on the circular profile rad z Distance of contact point from origin along the vane axis m l Distance of end of guide from origin along the vane axis m

1, 2

P P Fluid pressure Pa

Tvane Torque on one vane per unit width Nm/m

chamber

T Torque corresponding to one chamber (per unit width) Nm/m

total

T Total torque (instantaneous) Nm

C1 Contact point between rotor and stator profiles C2 Contact point between vane and the cam profile

 Angular position of C₁ rad

AVi Chamber area associated with vane i m²

dAVi Change in area of chamber associated with vane I m²



^{X Y,}



Vane coordinate system

X, Y

F F Force components vane coordinate system N

X, Y

R R Reaction force components in vane coordinate system N

x, y

R R Reaction force components in motor coordinate system N

m Gear module m

D Pitch circle diameter of gear m

Z Number of teeth

Z Difference in number of teeth on circular gear and flex gear

 Deflection of flex gear m

C, f

Z Z Number of teeth on circular gear and flex gear respectively

C, f

D D Pitch circle diameter of circular gear and flex gear respectively m

i, o

r r Inner radius and outer radius of approximated ring m

t Thickness of approximated ring m

F Deflecting force acting on the ring N

ds Small section of quarter ring m

d Angle subtended by dsat the centre rad

r,

F F_ Force component on the small section, ds, in radial and tangential direction respectively

N

M Bending moment at the small section ds ^Nm

U Total strain energy J

dU Strain energy on small section ds ^J

1, 2, 3

dU dU dU Component of dUdue to normal load, shear load and bending moment respectively

J C Correction factor for a rectangular cross-section in shear

M O Bending moment at free end Nm

A Cross-section area m²

E Elastic modulus Pa

G Shear modulus Pa

I Area moment of inertia m⁴

i, o

  Stress at inner fibre and outer fibre respectively Pa

i, o

c c Distance from neutral axis to inner fibre and outer fibre respectively (for curved beams with rectangular cross-section)

m e Distance from centroidal axis to neutral axis for curved beams with

rectangular cross-section

m r Radius of neutral axis for curved beam with rectangular cross-

section

m

r c Radius of centroidal axis for curved beam with rectangular cross- section

m

K a Stress concentration factor at gear root

, ; ,

Ji Jo Ki Ko

    Stress at inner and outer fibre at section J and K respectively Pa

_ , _

i mean o mean

  Mean stress at inner and outer fibre respectively Pa

_ , _

i amp o amp

  Amplitude of alternating stress at inner and outer fibre respectively Pa

Syt Material yield strength Pa

S e Material endurance limit Pa

f s Safety factor

r t Any radius on an involute gear profile m

t t Thickness of gear tooth at r _t m

d o Addendum diameter for external tooth OR dedendum diameter for internal tooth

m

b, b

d r Base circle diameter and radius respectively m

p Pressure angle rad

 ^cos1



rb ^/rt



rad

pt, pt

t r t and t r at pitch circle t m

dof Addendum circle diameter of flex gear m

dbf Base circle diameter of flex gear m

drf Root diameter of flex gear m

d rC Root diameter of circular gear m

dbC Base circle diameter of circular gear m

doC Addendum circle diameter of circular gear m

d C Mid-section circle diameter of flex gear m

P C Perimeter of mid-section circle of flex gear m

a e Major axis of ellipse m

P e Perimeter of ellipse m

e e Eccentricity

b e Minor axis of ellipse m



X Y j^, j



Coordinates of point on circular gear profile (flexgear)

j Angle subtended at the centre by arc between horizontal axis and



X Y j^, j



rad

C  _j  _j1 ^rad



XjC^,YjC



Point corresponding to



X Y on mid-section circle j^, j





^Xje,Yje



Point corresponding to



^XjC,YjC



on elliptical mid-section



^{X Yj, j}



Point on elliptical gear profile corresponding to



X Y j^, j



p c Length of arc bounded by two points on mid-section circle m '

p Length of elliptical arc m

e Angle subtended by elliptical arc at the centre rad

j Angle subtended at centre by elliptical arc bounded by horizontal axis and



Xje^,Yje



rad

t e Elliptical parameter in parametric equation of ellipse

C, C

X Y

  Vector difference of corresponding points on mid-section circle and elliptical mid-section

m u Number of pairs of meshing teeth supplied with fluid

1, 2, 3,....

A A A Area trapped between meshing pairs of teeth m²

r s Speed ration

vwg Speed of wave generator rpm

vf Speed for flex gear rpm



X_jm,Y_{j m}



Mid-point of line joining two successive points on elliptical gear profile

lj Distance between two successive points on elliptical gear profile m mj Slope of perpendicular to the line joining two successive points on

elliptical gear profile



X^jk^,Y^jk



Point of intersection of line perpendicular to line joining two successive on elliptical gear profile passing through the midpoint



X_jm,Y_{j m}



; and the perpendicular to it passing through the origin

dj Distance of



X^jk^,Y^jk



from origin m

Tj Torque contribution from a small region on elliptical gear profile on all four symmetric positions

Nm

Twg Torque on wave generator Nm

Tf Torque on flex gear Nm

L Length of flex gear cup m

t c Thickness of flex gear cup m

mean, alt

  Mean and alternating stress respectively Pa

Abbreviations

CAD Computer Aided Design

1-D One dimensional

TABLE OF CONTENTS

SAMMANFATTNING ... I ABSTRACT ... III FOREWORD ... V NOMENCLATURE ... VII TABLE OF CONTENTS ...XIII

INTRODUCTION ... 1

1.1 B

ACKGROUND

... 1

1.2 A

BOUT

A

TLAS

C

OPCO

... 2

1.3 P

ROBLEM

D

ESCRIPTION

... 2

1.4 P

URPOSE

... 3

1.5 D

ELIMITATIONS

... 3

FRAME OF REFERENCE ... 5

2.1 R

OTARY PERCUSSIVE ROCK DRILLING

... 5

2.2 R

OCK DRILL

(D

RIFTER

) ... 5

2.3 R

OTATION MECHANISM

... 6

2.4 H

YDRAULIC MOTORS

... 10

CONCEPTS ... 11

3.1 R

EQUIREMENT

S

PECIFICATION

... 11

3.2 C

ONCEPTS

G

ENERATED

... 11

3.3 C

ONCEPT EVALUATION

... 16

MULTI-CAM HYDRAULIC VANE MOTOR CONCEPT ... 19

4.1 W

ORKING

P

RINCIPLE

... 19

4.2 D

ESIGN CALCULATIONS

... 21

4.3 C

ONCEPT IMPROVEMENT

... 30

4.4 C

AM VANE MOTOR

: V

ANES ON ROTOR

... 32

4.5 C

AM VANE MOTOR

: V

ANES ON STATOR

... 47

4.6 D

IFFERENT NUMBER OF VANES

... 63

STRAIN WAVE GEAR HYDRAULIC MOTOR CONCEPT ... 69

5.1 W

ORKING

P

RINCIPLE

... 69

5.2 M

INIMUM NUMBER OF TEETH ON FLEXGEAR

... 70

5.3 M

ATHEMATICAL MODEL

... 76

5.4 M

OTOR ASSEMBLY DESIGN

... 85

DISCUSSION AND CONCLUSIONS... 91

6.1 D

ISCUSSION

... 91

6.2 C

ONCLUSIONS

... 92

FUTURE WORK ... 93

REFERENCES ... 95

APPENDIX A: INTEGRATION FOR AREA ENCLOSED ... 97

APPENDIX B: CALCULATION OF AREA COMPONENTS A

B

AND A

C

101

INTRODUCTION

This chapter provides a brief introduction, the background, the problem description, purpose and the delimitations of the thesis work.

1.1 Background

Blasthole drilling has been in practice since centuries for mining. To make better use of explosive force, miners started to place the explosives (gunpowder) in holes drilled into rock. One man drilling with the help of a drill steel and sledgehammer was an establish technology used in the eighteenth century (Figure 1). This physically demanding technology evolved slowly to give rise to powered drills (Atlas Copco Drilling Solutions LLC 2012). The need for drilling bigger and deeper holes into rocks at a faster rate led to the development of heavy and high power rock drills (e.g. Figure 2) mounted on movable drilling rigs, for e.g. the underground drilling rig XE4 from Atlas Copco (Figure 3).

Figure 1. Drilling using a drill steel and sledgehammer

Figure 2. A rock drill

Figure 3. Underground Drilling rig type XE4 for tunnelling.

1.2 About Atlas Copco

Atlas Copco is a world leading provider of sustainable productivity solutions. The group serves customers with innovative compressors, vacuum solutions and air treatment systems, construction and mining equipment, power tools and assembly systems. (Atlas Copco : Facts in brief n.d.) The Mining and Rock Excavation Technique business area provides equipment for drilling and rock excavation, a complete range of related consumables and service through a global network.

The business area innovates for sustainable productivity in surface and underground mining, infrastructure, civil works, well drilling and geotechnical applications. Principal product development and manufacturing units are located in Sweden, the United States, Canada, China and India. (Atlas Copco : Organisation- Mining Technique n.d.)

Atlas Copco Rock Drills AB in Örebro is a part of Mining and Rock Excavation Technique. The division is involved in development and manufacture of rock drilling machines, automation systems, tunnelling and mining equipment for various underground applications, procurement, warehousing and distribution of spare parts and shipping of finished drilling rigs, trucks and loaders to the sales companies and customers around the world. (Atlas Copco Rock Drills n.d.)

1.3 Problem Description

Before a part of tunnel can be blasted, a number of holes have to be drilled in a specific pattern for charging with explosives. Most of the drilling patterns have contour holes (Figure 4), which are fairly closely spaced in the roof and walls of the tunnel and charged with less powerful explosive.

By this means, crack formations in the side of the rock nearest to the tunnel surface are reduced and overbreak is avoided (Martin Lindfors 1985). It is advisable to choose a smaller diameter for the contour holes and drill them as close as possible to the edge of the tunnel.

Figure 4. Different holes drilled during tunnelling (drifting)

The current layout of the rotation mechanism of the rock drills involves a rotation motor which rotates an axle parallel to the shank. This rotation is transferred to the shank using a gear set (Figure 6). This layout makes the rock drill bulky. This limits the minimum distance at which the rock drill can be operated from the edge of the tunnel, therefore, limiting the minimum distance of the contour holes from the edge.

Therefore, a high emphasis is being laid at Atlas Copco for developing a rotation mechanism which would reduce the bulkiness of the rock drill.

1.4 Purpose

The goal of this thesis project was to evaluate the possibility to use the shank as a part of the rotation mechanism. The project involves development of new concepts that would address the problem and could replace the existing rotation mechanism.

The detailed goals of the project are as follows:

 Enlist requirement specifications for the rotation mechanism

 Develop concepts that would use the shank as a part of the rotation mechanism

 Shortlist concepts on the basis of requirement specifications

 Perform functional design calculations for the concepts

 Develop CAD models for the shortlisted concepts

The project was carried out for a particular model of rock drill. The concept can be later extended for use on different models of rock drills.

1.5 Delimitations

The design limitations state that no other sub-system of the rock drill (percussion, damping, feed and flushing) must be affected by the introduction of new concepts. It is advised to not increase the number of hoses and cables to the rock drill.

The mathematical models for the concepts have been developed only for ideal case, i.e. considering 100% efficiency. The torque and speed calculations have been performed for constant pressure difference and constant flow rate.

CAD models depicting the assembly layout of the motors were developed. The CAD model of the rock drill was not changed. Therefore, physical interfaces of the developed concept with the rock drill were not designed.

FRAME OF REFERENCE

This chapter provides information about percussive rock drilling, rock drills, rotation mechanism and hydraulic motors.

2.1 Rotary percussive rock drilling

The drilling principle is based upon the impact of a steel piece (piston) that hits a utensil (shank) which at the same time transmits the impact energy to the bottom of the blast hole by means of a final element called the bit.

The rotary percussive rigs are classified in two large groups, (Jimeno, Jimeno and Carcedo 1995)

 Top hammer: Rotation and percussion are produced outside the blast hole, and are transmitted by the shank adaptor and the drill steel to the drill bit.

 Down the hole hammer: Percussion is delivered directly to the drill bit, whereas rotation is performed outside the hole.

Percussive rock drilling involves four major processes; percussion, rotation, feed (thrust load) and flushing. (Figure 5)

Figure 5. Processes of percussive rock drilling (Jimeno, Jimeno and Carcedo 1995)

The four processes are described below, (Jimeno, Jimeno and Carcedo 1995)

1. Percussion: The impacts produced by repeated blows of the piston generate shock waves that are transmitted to the bit through the drill steel (in top hammer) or directly upon it (down the hole)

2. Rotation: Rotation of the drill string is required so that the inserts on the drill bit crush the rock in a new position for each percussion impact. This helps in uniform drilling and having good penetration rates.

3. Feed (thrust force): In order to maintain the contact of the drill bit with the rock, a thrust load or feed force is applied to the drill string.

4. Flushing: Flushing is the removal of broken rock (drill cuttings) from the blast hole by using pressurized air or water.

2.2 Rock drill (Drifter)

In modern drilling rigs, the four processes of percussion drilling are performed by drifters. A drifter is a hydraulic or pneumatic powered rock or ground drill placed on top of a feed which is like a rail on which the drill travels (drifts) on. Drilling using a drifter is known as drifting. The feed is attached with a flexible boom to a stationary or a mobile unit that contains the powerpack. (Drifter 2014)

The various sub-systems on a modern drifter rock drill from Atlas Copco (top hammer) are shown in Figure 6. The percussion system consists of a hydraulically actuated percussion piston which strikes against the shank adapter. The kinetic energy of the piston is transmitted, in the form of stress wave, via the shank adapter, drill steel and the drill bit to the rock, where it is used for crushing. The rotation is provided by a hydraulic motor, usually an orbital gerotor motor, which rotates the shank adapter through an axel and a gear train. The rotation torque needed to overcome the forces at the bit and the drill rod is also used to keep the threads on the drill string tightened.

The flushing air/water, introduced into the shank, is carried through a channel in the drill steel till the drill bit. Feed force is provided by moving the rock drill on the feed using a hydraulic feed cylinder (not in the picture).

Figure 6. Sub-systems of a rock drill

2.3 Rotation mechanism

As mentioned in the previous section, the shank rotation is provided by a hydraulic motor, usually an orbital gerotor motor (explained in 2.3.2), which rotates the shank adapter through an axel and a gear train.

The default rotation speed is such that the peripheral buttons on the drill bit are moved by a button diameter between each stroke. A high speed will lead to increased wear on the buttons, whereas a low speed would lead to abrasion as the buttons would lock into the already crushed rock. The rotation also provides sufficient torque to keep the threaded joints between drill rods tight.

(Grundström and Nordin 2007)

The required rotation speed (n_D) can be calculated for using the following formula, (Wijk 1995) 60 ^B

D

B

n z f

D

 (1)

Where D_Bis the bit diameter, f is the impact frequency in Hz and z_Bis the bit movement per blow.

Value of z_Bdepends on the type of button bit and the rock hardness. The recommended values are as follows,

 Button bits, normal rock : z_B 10 mm

 Button bits, very hard rock: z_B 8 mm

 Blade bits, normal rock: z_B 7.5 mm

For a 42 mm bit diameter (D_B), 65 Hz impact frequency ( f ) and button bits on normal rock ( 10 mm

zB  ), the required rotation speed would be 295.5 rpm.

2.3.1 Rotation hydraulics

The rotation system provides a constant flow when the directional valve is actuated. The pump has a constant speed. The simplified rotation hydraulic system is shown in Figure 7.

The pump is connected to the valve block which contains a pressure controlled valve and a directional valve that creates a constant flow to the motor. The directional valve is controlled by pilot pressures (XA and XB) set by the operator. The valve block also contains two pressure limiting valves that protect the system against extremely high pressure.

Figure 7. Rotation hydraulic system (Grundström and Nordin 2007)

The rock drill considered for this project comes with different sizes of rotation motor depending upon the application. The oil flow for rotation is,

 Maximum continuous: 75 litre/minute

 Maximum intermittent: 90 litre/minute

The 80 cc motor (smallest) provides a shank rotation speed of 300 rpm with approximately 52 litre/minute of oil flow.

2.3.2 Gerotor motor

A gerotor motor is a compact hydraulic motor that delivers high torque at low speeds. It consists of two units, an inner and outer rotor. The inner rotor is located off-axis and has one less tooth than the outer rotor. Each tooth of the inner rotor remains continuously in contact with the outer rotor. A gerotor can either have two fixed axes at an offset for each rotor (both the rotors spin) or a fixed outer rotor with the inner rotor revolving around the centre of the outer rotor while rotating about its own axis. The motor used in the rock drills are of the later type. The inner rotor has a trochoidal profile (formed by the locus of a point on a circle rolling on a circle). The outer rotor is formed by a circle and intersecting circular arcs.

The inner and outer rotors form sealed pockets of fluid. The pressurized fluid fed to the motor acts directly on the exposed inner rotor tooth via appropriate porting or a distributor valve. The inner rotor is thus caused to rotate relative to the stationary outer rotor.

The classification of gerotor motors based on the type of valve mechanism is as follows,

 Spool valve motors

 Disc valve motors

 VIS (valve-in-star) motors

The motor is shown in Figure 8. The outer rotor has rollers forming the teeth. Such a motor is also called a geroller.

Figure 8. Gerotor motor

The ports and distributor valve (for disc valve motor) are shown in Figure 9.

Figure 9. Ports and distributor valve

The off-axis rotation of the rotor is transmitted to the output shaft through a cardan shaft that has splined ends. The cardan shaft and the output shaft are shown in Figure 10.

Figure 10. Cardan shaft and output shaft

Gerotor motor produces high torque at low speed. The function diagram for MLHS 200 (disc valve motor with 200 cc displacement) is shown in Figure 11. The function diagram data is for average performance of randomly selected motors at back pressure 5+10 bar and oil with viscosity of 32 mm²/s at 50^oC. (M+S Hydraulic n.d.)

Figure 11. Function diagram for MLHS 200 motor from M+S Hydraulic (M+S Hydraulic n.d.)

2.4 Hydraulic motors

Hydraulic power transmissions are known to be compact for high power transmission. Figure 12 shows the most commonly acceptable regions of applicability for hydraulic, electrical and pneumatic power transmission media for rotation applications. It shows that hydraulic transmission media is better suited for high torque applications. (Hunt and Vaughan 1996)

Figure 12. Comparative operational regions (Hunt and Vaughan 1996)

Hydraulic motors commonly used are of the following types:

 Bent-axis axial piston motor

 Swash plate axial piston motor

 Vane motors

 Gear motors

 Lobe rotor motors (gerotor)

CONCEPTS

In this chapter a brief description of the concepts generated and their evaluation based on the requirement specification has been presented.

3.1 Requirement Specification

Requirement specification for the concepts have been listed in Table 1. The demanded specifications are necessary to be fulfilled by the generated concepts.

Table 1. Requirement specification

Requirement specification Requirement Demand/Wish

1 Space Radial diameter max. 150 mm Demand

2 High torque / low speed Displacement approx. 200 cc Demand 3 High Efficiency

4 Proper shock wave propagation No discontinuities in the shank Demand 5 Serviceability Shank easily removable /

replaceable

Demand 6 Independent system Independent rotation and

percussion

Demand 7 Rotation direction Both clockwise and anti-

clockwise

Demand 8 Number of hoses ≤ 2 (less than existing number of

hoses)

Demand 9 Integration Easy integration with the rock

drill

Wish

10 Simple design No complicated parts or systems Wish

11 Variable displacement Possible switch between high torque low speed to low torque high speed

Wish

The requirement for efficiency has not been specified because the efficiency calculations have not been performed in the project.

The motor (without any transmission) must be able to rotate the shank with a speed of 300 rpm with 60 litre/minute oil flow. Therefore, the displacement requirement was set to 200 cc.

The existing gear coaxial with the shank has an outer diameter of 147 mm. Therefore, the maximum allowable radial size was set to 150 mm so that the new concept could replace the gear in position without much changes required in the size of the casing.

3.2 Concepts Generated

Concepts for only hydraulic motors were generated because a pneumatic or electrical motor would require additional pneumatic or electrical lines running to the rock drill. A total of eight motor concepts were generated. The brief description of the concepts is given below.

3.2.1 Gerotor motor – With geared output shaft

The working of the motor would be same as a gerotor motor being used on the rock drill. The motor would surround the shank. The cardan shaft would be replaced by a geared output shaft

which would be supported by bearings on either side of the motor. The output shaft would be placed coaxially with the motor stator. The rotor would rotate the output shaft through gearing.

The hollow output shaft further rotates the shank which passes through it. (Figure 13)

Figure 13. Concept: Gerotor motor – With geared output shaft

The radial size of the motor would be high as there is unused radial space between the output shaft and the off-axis rotor. The displacement of the motor would be high and comparable to the existing motor design.

3.2.2 Gerotor motor – With shank passing through hollow cardan shaft

In this concept, the gerotor surrounds the shank. The shank passes through a hollow cardan shaft.

The cardan shaft transmits rotation from rotor to an output hub. The output hub further transmits rotation to the shank through the driver. (Figure 14)

Figure 14. Concept: Gerotor motor – With shank passing through hollow cardan shaft

The axial size of the motor would be high because of the long cardan shaft and the output hub. The displacement of the motor would be high and comparable to the existing motor design.

3.2.3 Axial piston motor – with swash plate

In this concept, the shank and the driver pass through a motor body. The working of the motor is same as a traditional axial piston motor. The high pressure hydraulic oil pushes the pistons out against a swash plate causing the swash plate (in case of a fixed cylinder drum) or the cylinder

drum (in case of fixed swash plate) to rotate. The shank would be rotated by the rotating part of the motor (swash plate or cylinder drum). (Figure 15)

Figure 15. Concept: Axial piston motor with swash plate

One complete stroke of piston would result in one revolution. Therefore, the displacement of the motor would be low and an additional transmission would be required. The motor would have low radial size and high axial size.

A mechanism to shut off fluid supply to select cylinders could be implemented to create a variable displacement motor.

3.2.4 Axial piston motor – with circular cam plate

The working of the concept is similar to the axial piston motor described in 3.2.3. The swash plate is replaced by a multi lobe circular cam plate. (Figure 16)

Figure 16.Concept: Axial piston motor with circular cam plate

This motor would have higher displacement than the motor with swash plate, because for one revolution of the motor the number of strokes of a piston required would depend upon the number

of lobes on the cam plate. Therefore, no additional transmission would be required. The motor would also have high axial size like the previous concept.

A mechanism to shut off fluid supply to select cylinders could be implemented to create a variable displacement motor.

3.2.5 Radial piston motor multi-lobe cam

In a radial piston motor, hydraulic pressure would move the pistons radially in the rotating unit.

This movement of the pistons against the cam surface would cause the rotation of the rotating unit.

The rotating unit would further rotate the shank through a driver. (Figure 17)

Figure 17. Concept: Radial piston motor multi-lobe cam

The motor would have a high displacement as the number of strokes of piston required for one revolution are dependent upon the number of lobes of the cam.

3.2.6 Multi-cam vane motor

In a multi-cam vane motor, the vanes are free to slide in guides on the rotation unit. The vanes are pressed against the profile of external cam. The volume trapped between two vanes is pressurized with hydraulic oil to cause the rotation of the rotating unit. (Figure 18)

Figure 18. Concept: Multi-cam vane motor

3.2.7 Strain wave hydraulic gear motor

A strain wave hydraulic gear motor concept is derived from the harmonic drive (Slatter and Slatter 2005) which allows high reduction ratios in mechanical drives. In the concept, the high pressure hydraulic oil is applied between the meshing teeth on one side of the wave generator major axis and on the diametrically opposite side. This would deform the flex gear resulting in rotation of the wave generator. A reduced rotation would be obtained on the flexgear through the strain wave gearing. The flex gear would rotate an output shaft which would further rotate the shank. (Figure 19)

Figure 19. Concept: Strain wave hydraulic gear motor

The motor is expected to exhibit very high displacement because of the high reduction ratio offered by the strain wave gearing. The efficiency of the motor might be poor because of difficult sealing of fluid chambers with high pressure fluid.

3.2.8 Screw motor

In a screw motor, the pressure of the hydraulic oil moving along the threads of the screws would cause the rotation of the screws. In the concept, the central screw would be hollow and the shank would pass through it. The central screw would rotate the shank through a driver. (Figure 20)

Figure 20. Concept: Screw Motor

The motor displacement would be high and an additional transmission would be required.

3.3 Concept evaluation

The concepts were evaluated for their compliance with the requirement specifications. Concepts were first evaluated with respect to the polar properties (yes/no). The concepts were supposed to satisfy all the mandatory properties. For the optional properties, an additional transmission was not desired, as it would complicate the design, and a variable displacement mechanism without complicating the design was desired (not mandatory). The evaluation for polar properties is shown in Table 2.

Table 2. Evaluation for polar properties

Properties

Concepts Gerotor

motor - with geared output shaft

Gerotor motor - hollow cardan shaft

Axial piston motor -

with swash

plate

Axial piston motor -

with circular

cam plate

Radial piston motor multi cam

Multi- cam vane motor

Strain wave hydraulic

gear motor

Screw motor

Mandatory

no shank

discontinuity Y Y Y Y Y Y Y Y

shank easily

removable Y Y Y Y Y Y Y Y

independent rotation and percussion

Y Y Y Y Y Y Y Y

rotation both

directions Y Y Y Y Y Y Y Y

free moving

shank Y Y Y Y Y Y Y Y

less or same number of hoses

Y Y Y Y Y Y Y Y

Optional

Transmission

required N N N Y N N N Y

Variable

displacement N N Y Y N N N N

The generated concepts were then evaluated using a weighted PUGH matrix. Each concept is compared to each other with respect to different factors as listed below:

 Radial size: A smaller radial size of the motor is desired. This property has high weightage because reducing the overall radial size of the rock drill is the main purpose of the project.

 Axial Size: A smaller axial size of the motor is desired. This property has a low weightage because a longer motor would only lead to slightly longer rock-drill which is not a major concern.

 Displacement: As the rotation mechanism is to be mainly used for high torque low speed application, a high displacement of the motor is desired.

 Efficiency: A higher efficiency of the motor is desired.

 Integration: The motor assembly must be easy to assemble and must also be easily mounted onto the rock drill.

 Simplicity: A simple design of the motor is desired.

 Robustness: The design must be robust.

The factors were assigned weights from 1 to 5. Each concept was scored from 1 to 5 for each factor. No baseline score was used as there is no existing design with motor unit surrounding the shank. The weighted PUGH matrix is shown in Table 3.

Table 3. Weighted PUGH matrix

Properties Weight

Concepts Gerotor

motor - with geared output shaft

Gerotor motor - hollow cardan shaft

Axial piston motor - with swash plate

Axial piston motor -

with circular

cam plate

Radial piston motor multi

cam

Multi- cam vane motor

Strain wave hydraulic

gear motor

Screw motor

Radial size 5 3 3 5 5 4 4 5 5

Axial size 2 5 2 2 2 5 5 5 1

displacement 4 4 4 1 4 4 5 5 1

efficiency 5 4 4 5 4 4 4 2 5

Integration 3 4 4 4 4 4 5 3 5

simplicity 3 4 4 5 4 3 4 4 4

robustness 4 4 4 5 5 5 5 4 5

weighted score 101 95 105 109 107 117 102 103

Based on the PUGH matrix, the multi-cam vane motor concept was chosen for further design.

Additionally, the strain wave hydraulic gear motor concept was also chosen for further design because of its novelty.

Both the concepts were designed at ideal conditions (100% efficiency) with constant pressure difference and oil flow rate. The pressure and flow parameters used are listed below:

 Input fluid pressure, P_in 120 bar1.2 10 Pa ⁷

 Return fluid pressure, P_out 1 bar 1 10 Pa⁵

 Fluid flow, Q60 l/min0.06 m /min³

MULTI-CAM HYDRAULIC VANE MOTOR CONCEPT

In this chapter the design of the multi-cam vane hydraulic motor, which is one of the selected concepts, is presented. The various improvements on the initial concept and the reasoning behind them are also discussed. The different versions of the concept have also been described.

4.1 Working Principle

The motor has a rotating cylindrical shaft (rotor) and a stationary housing (stator). The rotor also forms the output shaft that rotates the driver, which in turn rotates the shank adapter. The inner profile of the stator forms a cam ring. The volume chambers are separated by vanes which slide inside guides in the rotor. Alternatively, the rotor can have a cam ring profile and the vanes can be housed inside the stator. (Figure 21)

For design calculations, only the former system (Figure 21a) is considered.

a b

Figure 21. Concept versions, a) with vanes housed in rotor; b) with vanes housed in stator

The volume enclosed by the stator, rotor and the vanes is the active volume. The minimum and the maximum volumes are shown in Figure 22.

Figure 22. Maximum and minimum possible chamber volumes

High pressure hydraulic fluid through a port, whose position is fixed relative to the stator, would rotate the rotor such that the chamber volume increases from minimum to maximum. The direction of rotation can be reversed by swapping the pressure line with the return line. The vanes would regulate the opening and closing of fluid ports by uncovering and covering them respectively. The fluid port diameter must be smaller than the width of the vane. (Figure 23)

Figure 23. Operation by high pressure hydraulic fluid input and swapping of pressure ports with return ports to change direction of rotation

High pressure hydraulic fluid from the pressure line would be supplied to the vane guides so that the vanes are always pressed against the cam profile and the contact is never lost. (Figure 24)

Figure 24. High pressure hydraulic fluid used for maintaining contact between vane and cam profile

4.2 Design calculations

4.2.1 Stator cam profile

The profile of the cam ring is considered to be a sine profile because it can be easily represented as an equation which further simplifies the calculations. The profile in polar coordinates



^r,



can be given as follows,

sin sin

2 ^o 2 ^o

n n

ra ^   ^R  a ^ ^R

    (2)

Where R is the radius of the mean circle of the sine curve, _O a is the amplitude of the sine curve, n is twice the number of cams. Number of vanes ( )k was considered to be equal to n. Each vane was equally placed along the circumference of the rotor (angular spacing of 2 / n radians). The modification of

 

^ was made so that at an initial condition, when vane 1 is at  0, the volume enclosed by vane 1 and vane 2 2

0 n

 

   

 

  would be the minimum possible volume and the volume enclosed by vane 2 and vane 3 2 4

n n

 

   

 

 would be the maximum possible volume. The cam profile and the maximum/minimum chamber volume are shown in Figure 25.

Figure 25. Stator cam profile, position of vanes at initial condition and the maximum/minimum chamber volumes

A repetitive sub-cycle for rotation would be rotation of vane 1 from 0 to ² n

     and

corresponding rotation of vane 2 from ² to ⁴

n n

 

    . Therefore, the chamber volume enclosed by vane 1 and 2 would go from minimum to maximum and a rotor rotation of ²

n

 would

be obtained. The same sub-cycle occurs at 2

n locations for a rotation of ² n

 radians. Therefore,

for calculation of torque and speed, only one sub-cycle is analysed and then extended to the complete cycle.

Width of the stator and rotor unit is denoted by B.

4.2.2 Evaluation of points of contact

The considered cross section shape of the vane was rectangular (width W ) with circular head _p (radius r ) as shown in Figure 26a. The contact of the vane with the cam profile is shown in Figure _p 26b.

a b

Figure 26. a) Vane structure; b) contact geometry between vane and cam profile

In the figure,  is the actual angle of rotation of the vane from  0, ^{f x y}

 

^, is the equation of the cam profile in Cartesian coordinates,



XC^,YC



are the coordinates of the corresponding point of contact, i.e. C,



XR^,YR



are the coordinates of the centre of the circular part of the vane, i.e. R, and  is the angle of rotation of the contact point in the same coordinate system as used for . For a profile with equation ^{r  f}

 

^ in polar coordinates, the coordinates of the point on profile at angle ' would be,

   

( ') cos ' ; ( ') sin '

xr   yr   (3)

' '

cos( ') sin( '); sin( ') cos( ')

dx dr dy dr

r r

d d _{ }   d d _{ }  

 ^  _ ^  ^  _ ^

(4)

Slope

 

mn of the normal to the profile at any point is given by,

'

'

cos( ') sin( ') sin( ') cos( ')

n

dx dr r

dx d d

m dy dy dr

d d r

 

 

 

 

 

 







     



(5)

For the profile given in equation (2),

1 cos

2 2

dr n

d an 



 

    (6)

At an assumed angle, angle can be given by,

1

cos( ) sin( ) tan ( )

sin( ) cos( )

n

dr r

m d

dr r

d

 

 

 

 

 



 





  



(7)

Distance of the point of contact form the origin,R , _C

C ( )

R r  (8)

Coordinates of the point of contact



XC^,YC



,

   

cos ; sin

C C

X r_{ }  Y r_{ }  (9)

Coordinates of the centre of the circular head of the vane



XR^,YR



were calculated as,

   

p p

r cos ; r sin

R C R C

X  X   Y Y   (10)

The corresponding angle of rotation

 

^ of the vane was calculated as, tan 1 ^R

R

Y

  ^{ }X ^

  (11)

The distance of the centre of the circular part of the vane from the origin is given by,

2 2

R R R

R  X Y (12)

The angle

 

^ between the axis of the vane and the line joining the point of contact with the centre of the circular head of the vane (Figure 26b) is calculated as follows,

1

sin( ) cos( )

sin( ) cos

R p

R p

R r

R r

  

 ^  

 

  

   

(13)

 

2

     (14)

 

     (15)

The design parameters used for calculations are given in Table 4.

Table 4. Design parameters

Number of crests and troughsn 12

Number of vanesk 12

Amplitude of sine profilea 4 mm = 4 10 ^3 m Radius of vane headr _p 7 mm = 7 10 ^3 m Width of vaneW _P 7 mm = 7 10 ^3 m

Radius of rotor profileR_I



^{42  3a}



^{= 54 mm= 54 10 3 m}

Radius of mean circle of sine profileR _O

1 cos sin 1

2

p

I p

p

R r W a

r

    

 

     

= 59 mm

= 59 10 ^3 m

For 2 2

 

    , corresponding values of  were calculated. (Figure 27)

Figure 27. Corresponding values of parameters γ and α for φ = -90  90 deg

For further calculations, values of , R_C, R_R and were needed for any rotation angle  . These were obtained by first obtaining  value for any rotation angle  by reverse interpolation of tabular data of vs   using inbuilt Matlab function for 1-D interpolation by table lookup, interp1.

Spline interpolation method is used. This method uses not-a-knot end conditions. The interpolated value at a query point is based on a cubic interpolation of the values of neighbouring grid points in each respective dimension (Support: Documentation: interp1 n.d.). Then the R_C, R_R and  values were obtained for the calculated value of  using equations (6) to (15).

4.2.3 Speed calculation

As 2

k chambers (alternate chambers) would be pressurised simultaneously, change in volume of the pressurized chambers must be calculated to calculate the speed of the motor.

For volume calculation, the chamber bounded by vanes 1 and 2 was considered (Figure 25). For any rotation angle ( ), vane 1 would be at angular position,₁ , and vane 2 would be at angular

position ₂ ²

k

    . The chamber area was calculated by the following equation,

Chamber a b c

A A A A (16)

Aa is the area bounded by the cam profile and the circular inner profile for an angle between

1 2

   . A_b is the portion of A_a occupied by the vanes. A is the area that is added or removed _c from A_a because contact points between vane and the cam profile might not lie on the axis of vanes. (Figure 28)