Sage Journals: Discover world-class research

Abstract

The average time between failures of Chinese combine harvester was far below the advanced combine harvester in Europe and America. A load test system was designed and built on the transmission structure of combine harvester. The dynamic load of the transmission structure was tested on rice harvesting in the field and used to compare with calculated values of maximum load capacity. The damage mechanism and load distribution of transmission structure would be obtained. The test results showed that the average and peak power consumptions of each working part increased with the increase in the speed of the combine harvester. As the stubble height increased, the average and peak power consumptions of each component correspondingly increased with the increase in the load on the entire machine. The power consumptions of the left and right half shafts of the chassis gradually increased, whereas the power consumptions of the other components did not change significantly. The power consumption of the intermediate shaft ranged from 29.53 to 35.92 kW, and the average power consumption was 34.03 kW when harvesting was done at full load in the field. The limit power consumptions of the left and right half axles of the chassis were 9.35 and 11.55 kW.

Keywords

Rice harvesting combine harvesters load testing transmission system power distribution

Introduction

The transmission system of a combine harvester distributes and transmits the power generated by the engine to the working parts such as the walking system, header, conveying trough, threshing separation device, and cleaning device. During normal harvesting operations, the performance of its drive system was closely related to the harvesting efficiency, and the torque is the main load acting on the drive components of the combine harvester, the efficiency and lifespan of this type of harvester were low.¹

In recent years, extensive research had been conducted on combine harvesters under loads applied during maneuvering. Li et al.² developed the parametric modeling of the combine harvester Julong-280 chassis frame using the three-dimensional modeling software UG. After meshing, the modal parameters as well as the modal vibration shapes were calculated utilizing the NX Nastran. Based on the comparative analysis of the characteristics of the natural frequency and the external excitation frequency of the frame, the structural optimization scheme was proposed. Ning et al.³ summarized the research status of air and screen cleaning device of grain combine at home and abroad, and the development trend of air and screen cleaning device of grain combine is put forward. Heuler and Klatschke⁴ studied a linear cumulative fatigue damage mechanism of mechanical structures by testing the dynamic loads during crop harvesting. Stellmach et al.⁵ used a torque sensor to measure and evaluate the load of the walking chassis of a self-propelled forage harvester, and the stress signals were counted through the rainflow algorithm for damage calculation. Friesen et al.⁶ designed a control system that can adjust the machine walking speed to control its feed amount based on the drum torque and studied the relationship between the feed amount and the threshing loss rate. Pytka et al.⁷ built a test system that can be used under off-road conditions. The system included four-wheel dynamometers, a steering robot, a differential GPS system, an inertial platform, a noncontact vehicle speed sensor, and an on-board computer. Typical vehicle dynamic tests were conducted,⁸ and a high-precision tractor test system was developed to detect and measure the various performance parameters of tractors and implement systems. Sudajan et al.⁹ conducted an experimental study on the influence of the rotation speed and feed amount of the threshing drum of the combine harvester on grain damage. The grain threshing damage increased with the increase in the threshing drum rotation speed and feed amount.

The above results had showed that the dynamic load of a combine harvester during operation is an important factor affecting its reliability,¹⁰ and the dynamic load observed during actual operation was inconsistent with the test bench results. Combine harvester manufacturers had conducted numerous field experiments, including load tests on traveling half shafts during operation, providing a theoretical basis for improving their design. Wang et al.¹¹ experimentally obtained the temporally varying data of the instantaneous rotational speed for an aircraft wheel at the instant of its contact with the test platform, using a noncontact photoelectric testing method. Moreover, the kinetic relationship between the transient rotational speed of the aircraft’s wheel and the horizontal impact force was established.

In this study, a crawler-type combine harvester was used as a test prototype, and based on the advanced load testing technology established at home and abroad, a wireless load testing system for collecting load signals from the chassis and header of the combine harvester and a wired load testing system for collecting the load signals from each working part were constructed. The dynamic load of the transmission structure was tested on rice harvesting in the field and used to compare with calculated values of maximum load capacity, the damage mechanism, and load distribution of transmission structure obtained.

Drivetrain analysis

Transmission structure and parameters

Transmission system structure

A crawler-type combine harvester adopts a double-horizontal axial-flow threshing separation structure and a cleaning method involving a single centrifugal fan and a double-layer vibrating screen, which is suitable for harvesting crops such as rice and wheat. Combine harvester mainly includes a reel, cutter, feeding auger, conveying trough, No. 1 threshing drum, No. 2 threshing drum, fan, cleaning screen, crawler-type walking chassis, and other components, which was showed in Figures 1 and 2.

Figure 1.

Transmission diagram of the main working parts of a combine harvester.

Figure 2.

Schematic of the driving system of a crawler-type combine harvester.

A crawler-type combine harvester mainly realized power transmission through the transmission shaft, chain drive, and belt drive. A part of the power generated by the engine was transmitted to the hydraulic continuously variable transmission, which transmitted the power to the left and right drive axles of the walking chassis through the gearbox, and these axles were connected to the combine harvester. The left and right crawler belts help the machine to crawl; another part of the power generated by the engine was outputted through the intermediate shaft, and a part of the power of the intermediate shaft was transmitted to No. 1 and No. 2 threshing drums and then through the No. 1 threshing drum shaft. The trough, the main shaft of the inclined trough, transmits part of the power to the header to drive the reel, feeding auger, and cutter, while another part of the power of the intermediate shaft was transmitted to the cleaning centrifugal fan, horizontal feeding auger, double auger, vibration cleaning screen, etc.

Parameters of the main components of the transmission system

To drive the main working components of a crawler-type combine harvester, such as the header, inclined conveying trough, No. 1 threshing drum, No. 2 threshing drum, the intermediate shaft, and the walking chassis, the operational reliability of the existing crawler-type combine harvester was accurately analyzed and calculated. First, the design of the chain drive, drive shaft, and belt drive of an existing crawler-type combine harvester was calculated and checked, and a crawler-type combine harvester was obtained through calculation and statistics. The parameters of the main working parts of the harvester was presented in Table 1.

Table 1.

Parameters of the main components of crawler-type combine harvester.

Working part	Parameters	Values
Engine	Rated speed (r/min)	2700
Engine	Rated power (kW)	52
Mechanical gearbox	Input speed (r/min)	2234
Mechanical gearbox	Design power (kW)	19–34
Drive wheel	Speed (r/min)	Шblock 200
Drive wheel	Design power (kW)	19–34
Intermediate shaft	Speed (r/min)	1600
Intermediate shaft	Design power (kW)	25–35
No. 1 threshing drum	Speed (r/min)	760
No. 1 threshing drum	Design power (kW)	10–15
No. 2 threshing drum	Speed (r/min)	760
No. 2 threshing drum	Design power (kW)	10–15
Conveyor drive shaft	Speed (r/min)	340
Conveyor drive shaft	Design power (kW)	2–5
Reel	Speed (r/min)	35
Reel	Design power (kW)	1.0–1.5
Fan	Speed (r/min)	1850
Fan	Design power (kW)	0.5–1.0
Horizontal grain auger	Speed (r/min)	862
Horizontal grain auger	Design power (kW)	0.5–0.8
Combine harvester	Machine quality (t)	2.46

In the drive system of a combine harvester, the power distribution of each step was critical, including the cutting, feeding, conveying, threshing, separation, and cleaning of rice stalks. If any link fails, it would affect the harvest rate of the harvester. Therefore, when the combine harvester was working normally, all the links in the transmission system should bear the corresponding load. To ensure the operational reliability and trouble-free working of the crawler-type combine harvester in the field, it is necessary to calculate the load of each main transmission component acting on the harvester.

Belt strength calculation

The design criteria of the belt drive were as follows: under the condition that the belt drive does not slip, it had a certain fatigue strength and life, and the rated power $P_{r}$ of a single V-belt can be obtained from “Mechanical Design.”¹²

P_{r} = (P_{0} + Δ P_{0}) \cdot K_{α} \cdot K_{L}

(1)

In the formula, $Δ P_{0}$ was the increment of the rated power $P_{0}$ of a single V-belt when the transmission ratio was not equal to 1, kW; $K_{α}$ was the correction factor when the wrap angle was not equal to 180°; $K_{L}$ was the correction factor when the belt length was not equal to the specific belt length specified in the experiment.

Based on the transmission structure of the crawler-type combine harvester showed in Figure 1, the basic parameters of the main transmission belts of each link of the harvester were showed in Table 2.

Table 2.

Basic parameters of the belt drive.

Basic parameters	Transmission belt from trough main shaft to header drive shaft	Drive belt from auger shaft to reel spindle	Intermediate shaft to feed auger shaft drive belt
Belt type	B	B	B
Small pulley reference diameter (mm)	130	111	143
Small pulley speed (r/min)	430	180	1600
Gear ratio	1.26	2.2	1.86
Small pulley wrap angle $α$ (°)	160	170	180
Belt length (mm)	3550	1500	2500
Number of V-bands	2	1	2

Based on the basic parameters of the belt drive provided in Table 2, the basic rated power $P_{0}$ of a single ordinary V-belt of each belt drive, the rated power increment of a single ordinary V-belt $Δ P_{0}$ , and the correction coefficient $K_{α}$ when the wrap angle was not equal to 180° were checked. The correction factor $K_{L}$ of the ordinary V-belt and the belt length were considered equivalent. The results were substituted into equation (1) to calculate the $P_{r}$ values of each transmission belt. The calculation results were showed in Table 3.

Table 3.

Calculation results of the power of the main drive belt.

Calculated parameters	Trough main shaft to header drive shaft drive belt	Auger to reel spindle drive belt	Intermediate shaft to feed auger shaft drive belt
Rated power $P_{0}$ (kW)	2.18	1.00	6.24
Rated power increment $Δ P_{0}$ (kW)	0.08	0.06	0.45
Correction factor $K_{α}$	0.95	0.98	1
Belt length correction factor $K_{L}$	1.09	0.92	1.03
Calculated power rating $P_{r}$ (kW)	2.34	0.96	6.89
Total power rating P (kW)	4.68	0.96	13.78

Transmission chain strength calculation

The intermediate shaft to the input end of No. 1 threshing drum shaft, No. 1 threshing drum to the input end of No. 2 threshing drum shaft, the input end of the No. 1 threshing drum shaft to the input end of the main shaft on the conveying trough, and the output end of the header drive shaft to the header were checked. The rationality of the design of the four chain drives at the input end of the auger was determined, and the allowable power of the four chain drives was calculated. The basic parameters of the four chain drives were showed in Table 4.

Table 4.

Basic parameters of the chain drive.

Basic parameters	Chain from the intermediate shaft to the No. 1 threshing drum shaft	Chain from the No. 1 threshing drum shaft to the No. 2 threshing drum shaft	Chain from the shaft of the No. 1 threshing drum to the shaft of the conveying trough	Chain of header shaft to header auger
Chain model	12A	12A	10A	8A
Number of chain rows	1	1	1	1
Number of small sprocket teeth	16	16	17	23
Small sprocket speed (r/min)	1600	1600	760	430

The actual allowable power of the chain drive could be expressed as follows:

[P_{0}] = \frac{P_{0} K_{m}}{K_{Z}}

(2)

In the formula, $P_{0}$ was the rated power of the A series roller chain under specific conditions, kW; $K_{m}$ was the row number coefficient of the roller chain; $K_{Z}$ was the small sprocket tooth number coefficient. Based on the basic parameters provided in Table 4, the rated power $P_{0}$ of the A series roller chains of each drive chain under specific conditions, the row number coefficient $K_{m}$ of the roller chain, the tooth number coefficient $K_{Z}$ of the small sprocket, and other coefficients were determined and substituted into equation (2); and the rated power value $[P_{0}]$ of each transmission chain was calculated. The calculation results were showed in Table 5.

Table 5.

Calculation results of the allowable power of the chain drive.

Calculated parameters	Chain from the intermediate shaft to the first roller shaft	Chain from the first threshing shaft to the second drum shaft	Chain from the first threshing shaft to the conveying trough shaft	Chain of the header shaft to the header auger
$P_{0}$ (kW)	41.6	41.6	15.6	4.37
$K_{m}$	1	1	1	1
$K_{Z}$	1.66	1.66	1.54	1.1
Calculated $P_{0}$ (kW)	25.06	25.06	10.13	3.97
Design power (kW)	10–15	10–15	2–5	1.5–2.2

Drive shaft strength calculation

The stress acting on the header transmission shaft, inclined conveying trough main shaft, No. 1 threshing drum shaft, and No. 2 threshing drum shaft during operation was calculated. The driving shaft of each working part was mainly affected by the torque during rotation, resulting in a torsional shear stress. The formula for calculating the maximum shear stress was:

τ_{\max} = \frac{T}{W_{t}}

(3)

In the formula, T was the torque acting on the transmission shaft, N m; $W_{t}$ was the torsional section coefficient of the shaft section. From $W_{t} = I_{P} / R$ , the solid axis $W_{t}$ could be obtained using the following formula:

W_{t} = \frac{π D^{3}}{16}

(4)

In the formula, D was the cross-sectional diameter of the shaft. From equation (3), it could be found that:

τ_{\max} = \frac{T}{W_{t}} \leq [τ_{T}]

(5)

From this, the design torque T of the transmission shaft could be obtained as:

T \leq W_{t} [τ_{T}]

(6)

The Calculation results of the shear stress was showed in Table 6, including the maximum torque of each transmission shaft calculated using equation (6).

Table 6.

Calculation results of the shear stress of the main drive shaft of the combine harvester.

Transmission shaft	Material	D (mm)	$W_{t}$ ( $m m^{3}$ )	$τ_{T}$ (MPa)	Maximum torque T (N m)
Header drive shaft	45 steel	26	3515.2	45	158.18
Chute spindle	40Cr	28	4390.4	50	241.472
First threshing drum	35 steel	38	10,974.4	30	384.104
Second threshing drum	35 steel	38	10,974.4	30	384.104
Intermediate shaft	40Cr	32	6553.6	50	360.448
Left axle	20CrMnTi	40	12,800	52	665.60
Right axle	20CrMnTi	40	12,800	52	665.60

The maximum torque of each drive shaft and the rated speed was used to design for the corresponding working parts; The calculation results of the rated power consumption was showed in Table 7.

Table 7.

Calculation results of the rated power consumption of the drive shaft of the combine harvester.

Transmission shaft	Design maximum speed (rpm)	Calculated power consumption (kW)	Theoretical design power consumption (kW)
Header drive shaft	520	8.61	4.3–10.7
Chute spindle	430	10.12	2–5
First threshing drum	760	26.20	10–15
Second threshing drum	760	26.20	10–15
Intermediate shaft	1600	57.71	25–35
Left axle	200	13.94	19–34
Right axle	200	13.94	19–34

According to Table 7, we found that the power that could be carried by the left and right half shafts of the walking chassis of the combine harvester was lower than the theoretical design power consumption range. However, the driving shaft of the header, the main shaft of the conveying trough, the shafts of the No. 1 and No. 2 threshing drums, and the output intermediate shaft could carry significantly more power than the theoretically designed power consumption.

Load testing system construction

Test rice material

The tested rice (Oryza sativa L. rice cultivar “zhendao10”) was grown in Danyang of Jiangsu Province, China. The average moisture content of the stalks was 68%, the average grain moisture content was 24%. The complete parameter estimation procedure and the results of all tests were repeated five times to reduce the impact of random influence on choice of training and validation set. The average data were analyzed statistically using a factorial design in SPSS software (version 13.0, SPSS Inc., CA, USA), then average results were compared by least significant difference (LSD) post-hoc test at the 5% significance level (p < 0.05).

Dynamic load test method

In order to develop field dynamic load test system, HAD-CYB-807S disc torque sensor (Beijing Westzh M & E Technology Co., Ltd., China. measurement accuracy of 0.25% FS and frequency response time of 100 μs), HAD-CYB-803S shaft torque sensor (Beijing Westzh M & E Technology Co., Ltd., China. measurement accuracy of 0.25% FS and frequency response time of 100 μs), gage patch (BE120-4AA(11) resistance strain gauges119.9 ± 0.1 Ω and MF-G14 silicone), and hall speed sensor (CHE18-15N11-HZF710) were used to test the torque and speed signals. Installation location and installation results of torque and speed sensor were showed in Figure 3. Based on the DH-5905 wireless torque and speed signals test method of WiFi wireless technology¹³ and the DH-5902 wired torque and speed signals test method were showed in some studies.¹⁴

Figure 3.

Field dynamic load test system.

Resistance strain gage load test system of reel cutting device was showed in Figure 4. HAD-CYB-807S disc torque sensor load test system of threshing cylinder was showed in Figure 5. HAD-CYB-803S shaft torque sensor load test system of engineering transmission shaft was showed in Figure 6. Resistance strain gage load test system of vibrating and cleaning device was showed in Figure 7. Resistance strain gage load test system of chassis drive axle was showed in Figure 8.

Figure 4.

Resistance strain gage load test system of transmission shaft.

Figure 5.

Disc torque sensor load test system of transmission shaft.

Figure 6.

Shaft torque sensor load test system of transmission shaft.

Figure 7.

Resistance strain gage load test system of sprocket wheel.

Figure 8.

Resistance strain gage of chassis drive axle with DH-5905 wireless signals acquisition system.

All the torque and speed signals were collected by the DH-5902 wired or DH-5905 wireless torque and speed signals acquisition system.^15,16 The torque and speed signals were sent to a laptop by a wireless card.

Calibration of wireless load test components

In order to establish the correspondence between the torque transmitted during the operation of the crawler combine harvester and the output voltage signal, the torque measurement module needed to be calibrated. The Equation of the calibration theory was showed in formula (7).

T = k \times V + b

(7)

Where, T was the torque value; V was the bridge output voltage.

Because the structure of the combine harvester was asymmetrical, the load applied to the left and right axles was different. The left and right half axles needed to be loaded and unloaded three times respectively. Standard torque was given by standard weights, at the same time the voltage output values were measured. The weight was gradually changed to obtain a corresponding set of voltage values. Finally, using Excel software to fit the calibration data, the trend line and calibration equation were obtained. The fitting straight line equation of the left and right half axle forward rotation calibration is showed in Figure 9.

Figure 9.

Forward rotation calibration equation of half axle: (a) half axle torque calibration test scene photo, (b) forward loading calibration equation of the left half axle, and (c) forward loading calibration equation of the right half axle.

Field load test on the combine harvester

Through the calculation of the main transmission system of the crawler-type combine harvester, the power that the belt, chain, and shaft of the main transmission parts could carry was obtained. However, the actual load carried by each main working part of the combine harvester when harvesting in the field and the corresponding power consumption and maximum limit power that needed to be carried under special working conditions were unknown. The power consumption under harvesting conditions was important for the design of the transmission parts of the harvester. For this purpose, the power consumptions of the header, conveying trough, and engine were selected. The power consumptions of the intermediate shaft, first and second threshing cylinders, cleaning system, and left/right semi-axis of the walking chassis were used as the main measurement parameters to conduct field load tests. The tests include load influencing factor tests and full load tests. In the former, a field comparative test was conducted on the forward speed, stubble height, and grain tank load, which affected the load distribution of the combine harvester. The power consumptions of the main working parts of the machine at full load and the limit power consumption under special working conditions were calculated.

Test and analysis of load influencing factors

When the combine harvester was being operated in the field, the loads of each working part were affected by the forward speed of the harvester (feeding amount), the stubble height (grass-to-grain ratio), and the condition of the walking surface. To test the main factors affecting the load of the combine harvester, referred to “General Provisions on Methods for Determining Test Conditions of Agricultural Machinery” (GB/T5262-2008), this study set the following test conditions under various combined harvesting conditions to collect torque and speed, such as low speed (0.45 m/s), medium speed (0.66 m/s), high speed (0.98 m/s), full grain box (1.1 t), half grain box (0.62 t), empty grain tank (0.14 t), low stubble (20 cm), medium stubble (30 cm), and high stubble (40 cm). The corresponding power consumptions were then calculated. The average moisture content of the rice grains was 24%, the moisture content of the stems was 72%, and the moisture content of the soil was 30%; each group of experiments was repeated twice during the test, and the average value was taken.

Walking speed test

When the combine harvester operated in the field, the harvesting length corresponding to field walking was 25 m, and the forward speed was calculated by recording the walking time within the range of 25 m. When harvesting in the field, three speed settings, namely low speed (0.45 m/s), medium speed (0.66 m/s), and high speed (0.98 m/s), were used for harvesting. The state of the granary was empty, and the stubble height was low. The torque and speed signals of the header, conveying trough, engine intermediate shaft, No. 1 threshing drum, No. 2 threshing drum, and the left and right half shafts of the walking chassis were collected. The corresponding power consumption values were computed.

The values of the intermediate shaft, conveying trough, No. 1 threshing drum, No. 2 threshing drum, and cleaning system were collected by the wired testing system when the crawler-type combine harvester was being operated in the field, as well as the values of the header and the left half of the chassis collected by the wireless load testing system. The output voltage signals of the shaft and the right half shaft of the chassis were substituted into the calibration equation to calculate the average loads at different forward speeds, as showed in Table 8.

Table 8.

Effect of forward speed on the average load (kW).

Component	Low speed	Medium speed	High speed
Intermediate shaft	24.88	30.53	34.92
Header	0.74	1.31	1.63
Conveying trough	0.39	1.08	1.78
No. 1 threshing drum	5.39	5.72	6.13
No. 2 threshing drum	9.10	9.56	10.55
Cleaning system	5.33	7.25	8.37
Left axle	0.73	1.68	2.14
Right axle	0.75	1.86	2.19

As showed in Table 8 the forward speed influenced the intermediate shaft, header, conveying chute, No. 1 threshing drum, No. 2 threshing drum, cleaning system, and left and right half shafts of the walking chassis. As the forward speed of the combine harvester increased, the average and peak power consumptions of each working part increased accordingly.

Effect test of stubble height

To study the effect of the stubble height (grass-to-grain ratio) on the load of the combine harvester during field harvesting, the forward speed was set as the medium speed, the granary was empty, and the load test was conducted under three harvesting conditions, such as low stubble (20 mm), medium stubble (30 mm), and high stubble (40 mm). The different stubble heights during field harvesting were showed in Figure 10.

Figure 10.

Different rice stubble heights: (a) low stubble, (b) medium stubble, and (c) high stubble.

The collected output voltage signals of the intermediate shaft, header, conveying chute, No. 1 threshing drum, No. 1 threshing drum, cleaning system, and chassis left half/right walking half shafts were substituted into the calibration equation to calculate the corresponding average loads, as showed in Table 9.

Table 9.

Effect of different stubble heights on the average load (kW).

Component	Low stubble	Medium stubble	High stubble
Intermediate shaft	33.91	29.26	25.59
Header	1.39	1.08	0.75
Conveying chute	1.09	0.93	0.69
No. 1 threshing drum	5.75	4.03	3.79
No. 2 threshing drum	13.35	11.86	8.22
Cleaning system	7.33	7.25	6.72
Left axis	1.78	1.76	1.73
Right axis	1.92	1.82	1.75

From Table 9, with the increase in the stubble height, the peak power consumptions of the intermediate shaft, header, conveying chute, No. 1 threshing drum, No. 2 threshing drum, and left and right half shafts of the chassis decreased.

Machine load impact test

To study the effect of the load of the entire machine (the weight of the grain in the grain box) on the load of the combine harvester in the field, three field harvests of rice were conducted under different load conditions, such as empty bin (0.14 t), half bin (0.62 t), and full bin (1.1 t). The collected output voltage signals of the intermediate shaft, header, conveying chute, No. 1 threshing drum, No. 2 threshing drum, cleaning system, and left and right half shafts of the chassis were substituted into the calibration equation to calculate the different machine loads. The calculated average loads was showed in Table 10.

Table 10.

Effect of different machine loads on the average load (kW).

Component	Low speed	Medium speed	High speed
Intermediate shaft	30.91	31.37	31.54
Header	1.46	1.42	1.39
Conveying chute	1.19	1.06	0.96
No. 1 threshing drum	5.75	5.79	5.62
No. 1 threshing drum	10.35	10.39	9.06
Cleaning system	7.68	7.43	7.34
Left axle	1.73	2.08	2.35
Right axle	1.75	2.26	2.76

As showed in Table 10, as the grain stored in the grain tank increased in volume (the load of the combine harvester increases), under the same working conditions, the power consumptions of the intermediate shaft, header, conveying trough, No. 1 threshing drum, and No. 2 threshing drum increased whereas the power consumption of the cleaning system did not change significantly. However, with the increase in the load of the entire machine, the power consumptions of the left and right half shafts of the chassis gradually increased.

Road condition impact test

A fast-running chassis load test was conducted under four driving conditions (sandy road, cement road, field soft mud road, and dry mud road). The chassis load test surface of the combine harvester was showed in Figure 11(a) when moving straight ahead on a gravel road. The chassis load test surface of the combine harvester was showed in Figure 11(b) when moving quickly on a cement road.

Figure 11.

Test road conditions: (a) gravel pavement, (b) cement pavement, (c) soft mud pavement, and (d) dry dirt pavement.

The chassis load test ground was showed in Figure 11(c) when the combine harvester was fully loaded and was running on a soft mud road in the field. The chassis load test ground was showed in Figure 11(d) when the combine harvester was moving on a dry dirt road.

The collected output voltage signals of the left and right axle shafts of the chassis were substituted into the calibration equation to calculate the average loads at full load, as showed in Table 11.

Table 11.

Influence of different pavements on the average loads of the left and right half axles (kW).

Component	Gravel pavement	Cement pavement	Soft mud pavement	Dry dirt pavement
Left axle	1.48	1.70	2.47	2.55
Right axle	1.73	2.49	2.58	2.83

From Table 11, it could be found that the power required by the combine harvester was the highest when moving on the soft mud road, and the power consumption of the right half shaft was slightly greater than that of the left half shaft.

Tests under different walking speed

The length of the test was 25 m, and the granary was empty. The change of harvester speed was obtained by controlling the gear position. At the same time, the torque and speed signals of the left and right axles of the walking chassis were collected. After data processing, the relationship between power consumption and speed was obtained. And the result was showed in Table 12. Effect of different forward speeds was showed in Figure 12.The data bar distribution diagram was showed in Figure 13.

Table 12.

Power consumption of axles.

Axle of test	Velocity (m/s)
	0.45	0.66	0.99
	Power consumption (kW)
Left axle	0.73	1.68	2.14
Right axle	0.75	1.86	2.19

Figure 12.

Effect of different forward speeds.

Figure 13.

The data bar distribution diagram.

When the speed of the combine harvester was increased, the power transmitted by the half axle was increased accordingly. The power consumption of left and right axles increased correspondingly, and they were approximately equal.

By comparing the data in Figure 13, the results were that: (1) when the slope angle was increased, the torque of axles would increase under the tested conditions. The increase of the slope angle of the climbing test would result in the average torque of the left and right axles increasing by approximately 10%. In the parking and emergency stop experiments, the increase of the slope angle would increase the average torque of the left axle by about 30%, and the average torque of the right axle would increase by about 45%. (2) When the slope angle was the same, taking the slope B as an example, the average working torque of the climbing slope test of the left and right half axles was increased by about 20% compared with the parking and the emergency stop condition. And the load of parking condition is similar to the emergency stop condition. (3) Under the three conditions, the load on the right half axles was greater than the load on the left one.

Ultimate load test of each component

Harvest ultimate load test

When the combine harvester was harvesting in the field, each working part was subjected to different speed (low speed (0.45 m/s), medium speed (0.66 m/s), and high speed (0.98 m/s)) and full grain box (1.1 t) loading. The load test under different combination conditions of half grain tank (0.62 t), empty grain tank (0.14 t), low stubble (20 cm), medium stubble (30 cm), and high stubble (40 cm) showed that the high speed, full-load, and low-stubble harvesting condition required the highest load. Therefore, five full-load, high-speed, low-stubble harvesting tests were conducted at intervals to test the normality of the intermediate shaft, header, No. 1 threshing drum, No. 2 threshing drum, and cleaning system. The load and peak load at the time of harvest and the average test results are showed in Table 13.

Table 13.

Average power consumption of the components of the combine harvester when working at full load (kW).

Tested part	Normal power consumption range at full load (kW)					Average value
Tested part	1	2	3	4	5	Average value
Intermediate shaft	29.53	33.91	34.88	35.92	35.90	34.03
Header	1.39	1.63	1.31	1.46	1.46	1.45
Conveying chute	1.03	0.94	1.08	1.06	1.45	1.11
No. 1 threshing drum	6.13	5.39	6.04	5.72	5.62	5.78
No. 2 threshing drum	10.65	9.10	10.45	9.06	10.84	10.01
Cleaning system	6.33	6.72	6.26	6.81	6.87	6.60

The power consumption distribution of the crawler-type combine harvester when harvesting at full load in the field was showed in Table 13.The power consumption range of the intermediate shaft was 29.53–35.92 kW, and the average power consumption was 34.03 kW. The power consumption range of the header was 1.31–1.63 kW, and the average power consumption was 1.45 kW. The power consumption range of the conveyor trough was 0.94–1.45 kW, and the average power consumption was 1.11 kW. The power consumption range of the first drum was 5.62–6.13 kW, and the average power consumption was 5.78 kW. The power consumption range of the drum was 9.06–10.84 kW, with an average power consumption of 10.01 kW. The power consumption range of the cleaning system was 6.33–6.87 kW, with an average power consumption of 6.60 kW.

Chassis ultimate load test

The average and ultimate loads required by the left and right drive shafts of the chassis were tested when the chassis of the combine harvester was under full and empty load conditions and under low-speed conditions, medium-speed conditions, and high-speed conditions. The crawler-type combine harvester was fully loaded on a 13° ramp (as showed in Figure 14), fully loaded and turned on dry ground (as showed in Figure 15), and quickly moved straight on a muddy ground. Load tests were conducted on sandy roads, cement roads, etc.

Figure 14.

Climbing a 13° slope backwards at full load.

Figure 15.

Turning in place at full load.

The average power consumption tested results of the chassis when the crawler-type combine harvester was fully loaded were showed in Table 14.

Table 14.

Average power consumption test results of the chassis (kW).

Working conditions	Left axle	Right axle
Low speed, empty	0.73	0.75
Medium speed, empty	1.78	1.92
High speed, empty	2.14	2.19
Low speed, full	0.97	1.14
Medium speed, full	2.26	2.67
High speed, full	3.61	4.85
Almost full, dry soil	3.43	4.55
Almost full, cement road	1.80	2.35
Slow full, climb 13	1.98	2.19
Turn left on dirt road	0	2.12
Turn right on dirt road	2.23	0
Sudden breaking at full load	1.68	1.96

Transmission system calibration of combine harvester

The power of the engine on the combine harvester was mainly distributed to the working parts such as the walking chassis, header, conveying trough, threshing and separating drums, and cleaning device. To withstand the extreme working conditions of each working part, it was necessary to check the design value and ultimate load of the main transmission parts and compare whether the main transmission parts meet the requirements.

The dynamic loads of working parts were tested at forward speed of 0.98 m/s and stubble height of 20 cm. The torque signals of every transmission shaft were recorded in laptops. The torque signals were showed in Figure 16. The rotation speeds of each transmission shaft were in a fluctuation range of ±10 rpm.

Figure 16.

Dynamic torque signals of working parts.

Strength evaluation of the transmission chain of combine harvester

To compare the reliability of the chain transmission of the main transmission components, the parameter values of the intermediate shaft to the input end of No. 1 threshing drum shaft, the intermediate shaft to the input end of No. 2 threshing drum shaft, and the input end of the No. 1 threshing drum shaft were calculated. The allowable powers of the four chain drives from the end to the input end of the main shaft of the conveying trough and the output end of the header drive shaft to the input end of the header auger were compared with the test results. The comparison results were showed in Table 15.

Table 15.

Comparison between calculated and measured allowable loads of the chain drive.

Transmission position	Intermediate shaft to the first threshing drum shaft	No. 1 threshing drum to the second threshing drum shaft	No. 1 threshing drum shaft to the conveyor trough	Header shaft to the header auger shaft
$P_{0} / kW$	41.6	41.6	15.6	4.37
$K_{m}$	1	1	1	1
$K_{Z}$	1.66	1.66	1.54	1.10
Calculated $P_{0}$ (kW)	25.06	25.06	10.13	3.97
Design power consumption (kW)	10–15	10–15	2–5	1.5–2.2
Measured average value (kW)	18.35	10.01	2.56	0.76
Peak power consumption (kW)	36.17	28.52	10.20	2.42

Transmission belt strength evaluation of combine harvester

The allowable powers of the transmission belt between the main shaft of the conveying trough and the transmission shaft of the header, the transmission belt between the auger and the tube shaft of the reel, and the transmission belt between the intermediate shaft and the auger shaft of the fan were compared, as showed in Table 16.

Table 16.

Comparison between calculated and measured allowable loads of the belt drive.

Transmission position	Main shaft on conveying chute to the header drive shaft	Auger shaft to reel shaft	Intermediate shaft to fan auger shaft
$P_{0} / kW$	2.18	1.00	6.24
$Δ P_{0} / kW$	0.08	0.06	0.45
$K_{α}$	0.95	0.98	1
$K_{L}$	1.09	0.92	1.03
$P_{r} / kW$	2.34	0.96	6.89
$P / kW$	4.68	0.96	13.78
Measured mean (kW)	1.45	0.24	6.60
Peak value (kW)	4.87	0.46	21.63

Optimization of the transmission system of combine harvester

Through the load calculations of the header, conveying trough, engine intermediate shaft, No. 1 threshing drum, No. 2 threshing drum, cleaning system, belt, shaft, and chain in the left/right half shaft transmission system of the walking chassis and a comparison with the average and limit power consumptions under full-load harvesting measured in the field, it was found that some of the belts could not meet the limit requirements for harvesting, the power consumption of some of the chains was too low to meet the normal harvesting requirements, and some of the axes could not meet the extreme requirements of harvesting. The belt, shaft, and chain of the combine harvester should be improved on the basis of the field load test results.

Improvement and optimization of the drive belt of combine harvester

From the verification of the transmission system in the fourth section, it could be concluded that the belt between the auger and the shaft of the reel and that between the intermediate shaft and the auger shaft of the fan could be improved. As showed in Table 17.

(1) The actual full-load average power consumption of the belt between the auger and the reel tube shaft was 0.24 kW, the maximum peak value was 0.46 kW, and the design power consumption of the belt was 0.96 kW. Because the power consumption of the reel is stable and the fluctuation frequency was low, the belt here could be changed to an A-type belt; the basic rated power $P_{0} = 0.64 kW$ , rated power increment $Δ P_{0} = 0.03 kW$ , wrap angle correction coefficient $K_{α} = 0.98$ , belt length correction coefficient $K_{L} = 0.92$ , and p = 0.6 kW, which meet the requirements.

(2) The actual full-load average power consumption of the belt between the intermediate shaft and the fan auger shaft was 6.60 kW, the maximum peak value was 21.63 kW, and the design power consumption of the belt was 13.78 kW. The belt here could not meet the peak power consumption requirements. The C-type double-belt could be chosen, with basic rated power $P_{0} = 12.63 kW$ , rated power increment $Δ P_{0} = 1.25 kW$ , wrap angle correction factor $K_{α} = 1$ , and belt length correction factor $K_{L} = 0.93$ , $P_{r} = 12.91 kW$ , p = 25.82 kW, which meet the requirements.

Table 17.

Basic parameters of the belt drive to be improved.

Transmission position	Auger to the reel shaft	Intermediate shaft to fan auger shaft
Original design power consumption (kW)	0.96	13.78
Measured mean (kW)	0.24	6.60
Limit power consumption (kW)	0.46	21.63
Center distance (mm)	435	845
Small pulley reference diameter (mm)	111	143
Small pulley speed (r/min)	180	1600
Transmission ratio	2.20	1.86

Improvement and optimization of the drive chain of combine harvester

The design of the chain between the intermediate shaft and the shaft of the No. 1 threshing drum and the chain between the shafts of the No. 2 threshing drum was unreasonable. As showed in Table 18 .

Table 18.

Basic parameters of the chain drive to be improved.

Transmission position	Intermediate shaft to No. 1 threshing drum shaft	No. 1 threshing drum shaft to No. 2 threshing drum shaft
Original design power consumption (kW)	25.06	25.06
Measured average value (kW)	18.35	10.01
Limit power consumption (kW)	36.17	28.52
Driving wheel speed (r $mi n^{- 1}$ )	1600	1600
Driven wheel speed (r $mi n^{- 1}$ )	760	760

The chain between the intermediate shaft and the shaft of the No. 1 threshing drum could withstand a power of 25.06 kW. The measured full-load normal working power consumption was 18.35 kW, and the peak power consumption was 36.17 kW. The power that the chain between the No. 1 threshing drum and the No. 2 threshing drum shaft could withstand is 25.06 kW, the measured full-load normal working power consumption was 10.01 kW, and the peak power consumption was 28.52 kW.

(1) A single row of 16A chains was used between the intermediate shaft and the No. 1 threshing drum shaft: $P_{0} = 62.4 kW$ , $K_{m} = 1$ , $[P_{0}] = 35.79 kW$ , meeting the transmission requirements.

(2) Since the No. 1 and No. 2 threshing drum shafts and the intermediate shaft and the No. 1 threshing drum shaft were connected via chain links, the selection of the chain should follow the standard of the chain between the intermediate shaft and the No. 1 threshing drum shaft.

Improvement and optimization of the drive shaft of combine harvester

Based on the verification of the transmission shaft in Section “Field load test on the combine harvester,” the header transmission shaft and No. 2 threshing drum shaft can be improved. The main parameters were showed in Table 19.

Table 19.

Main parameters of shaft to be improved.

Transmission shaft	Header drive shaft	No. 2 threshing drum shaft
Material	45# steel	35# steel
Minimum section diameter (mm)	26	38
Calculated power consumption (kW)	8.61	26.20
Original design power (kW)	4.3–10.7	10–15
Measured average value (kW)	1.45	10.01
Peak power (kW)	4.87	28.52

The following improvements can be made to the header drive shaft and No. 2 threshing drum shaft:

(1) Considering the economy, the diameter of the header drive shaft could be appropriately reduced, or other materials, such as 35# steel, can be selected.

(2) For the No. 2 threshing drum shaft, the diameter of the shaft could be appropriately increased, or 45# steel can be used.

Power distribution of each working parts in rice harvesting could be obtained and showed in Figure 17. Peak power of each working parts in rice harvesting could be obtained and showed in Figure 18.

Figure 17.

Power distribution of each working parts in rice harvesting.

Figure 18.

Peak power of each working parts in rice harvesting.

Figures 17 and 18, the material maximum power of reel cutting table was 8.61 kW, but the normal dynamic load power was 1.45 kW, the Peak power was 4.87 kW, the strength of reel cutting table drive shaft was enough. The material maximum power of Conveyor Bridge was 10.12 kW, but the normal dynamic load power was 1.11 kW, the Peak power was 6.38 kW, the strength of reel cutting table drive shaft was enough. The material maximum power of I-threshing cylinder 26.20 kW, but the normal dynamic load power was 5.78 kW, the Peak power was 23.60 kW, and the strength of reel cutting table drive shaft was enough. The material maximum power of II-threshing cylinder was 26.20 kW, but the normal dynamic load power was 10.01 kW, the Peak power was 28.52 kW, the strength of reel cutting table drive shaft was not enough. The material maximum power of Intermediate shaft of engine was 57.71 kW, but the normal dynamic load power was 34.03 kW, the Peak power was 49.62 kW, the strength of reel cutting table drive shaft was enough. The material maximum power of left drive shaft of chassis was 13.94 kW, but the normal dynamic load power was 3.61 kW, the Peak power was 9.35 kW, and the strength of reel cutting table drive shaft was enough. The material maximum power of right drive shaft of chassis was 13.94 kW, but the normal dynamic load power was 4.85 kW, the Peak power was 11.55 kW, the strength of reel cutting table drive shaft was enough.

Conclusions

In this study, a load testing system for the chassis, header, and various working parts of a crawler-type combine harvester was constructed, and the torque and rotation speed under different working conditions were tested. The load changes under different working conditions in terms of the various field load influencing factors were analyzed. The following conclusions can be drawn:

(1) Through a load test and a comparison test, the power consumptions of the working parts of the combine harvester were found to be the highest under harvesting conditions characterized by high speed, full-load, and low stubble height. The power consumption of the shaft was slightly higher than that of the left half shaft.

(2) The load testing system constructed for the combine harvester has a reasonable installation structure, is stable, reliable, and can accurately obtain the working loads of each working part of the harvester.

Footnotes

Handling Editor: Chenhui Liang

Declaration of conflicting interests

The author(s) declared no potential conflicts of interest with respect to the research, authorship, and/or publication of this article.

Funding

The author(s) disclosed receipt of the following financial support for the research, authorship, and/or publication of this article: This work was supported by the Jiangsu Agriculture & Animal Husbandry Science and Technology Vocational College (NSF2022ZR14) and Taizhou Science and Technology Support Plan (Agriculture) Project (TN202208, TN202219).

ORCID iD

Guoqiang Wang

Data availability

The data used to support the findings of this study can be obtained from the corresponding author upon request.

References

Špokas

Steponavičius

Žebrauskas

, et al. Reduction in adverse environmental impacts associated with the operation of combine harvesters during the harvesting of winter oilseed rape. J Environ Eng Landsc Manag 2019; 27: 72–81.

Sun

Pang

, et al. Finite element mode analysis and experiment of combine harvester chassis. Trans Chin Soc Agric Eng 2013; 29: 38–46.

Ning

Jin

Yin

, et al. Research status and development trend of air-and-screen cleaning device for cereal combine harvesters. J Chin Agric Mech 2018; 39: 5–10.

Heuler

Klatschke

Generation and use of standardised load spectra and load–time histories. Int J Fatigue 2005; 27: 974–990.

Stellmach

Braun

Wächter

, et al. On load assumptions for self-propelled forage harvesters. Int J Fatigue 2021; 147: 106114.

Friesen

Zoerb

Bigsby

FW.

Controlling federates automatically. Agric Eng 1966; 47: 434.

Pytka

Tarkowski

Fijałkowski

, et al. An instrumented vehicle for offroad dynamics testing. J Terramechanics 2011; 48: 384–395.

Singh

RC.

Computerized instrumentation system for monitoring the tractor performance in the field. J Terramechanics 2011; 48: 333–338.

Sudajan

Salokhe

Triratanasirichai

PM—power and machinery. Biosyst Eng 2002; 83: 413–421.

10.

Tang

Wang

, et al. Experimental study of multi-source dynamic load of combine harvester in rice harvesting. Int J Agric Eng 2016; 25: 46–52.

11.

Wang

, et al. A method for determining the horizontal impact load based on the rotational speed of the aircraft’s wheel in a landing gear drop test. Int J Crashworthiness 2018; 23: 627–634.

12.

. Mechanical design. Beijing: Higher Education Press, 2019.

13.

De Angelis

Pecce

. Model assessment of a bridge by load and dynamic tests. Eng Struct 2023; 275: 115282.

14.

, et al. Construction and performance experiment of load test system for half axle of combine harvester. Trans Chin Soc Agric Eng 2013; 29: 35–41.

15.

Golinelli

Spaggiari

Castagnetti

, et al. Development of a driveshaft torque transducer for low-cost structural health monitoring of off-highway vehicles. Proc IMechE, Part L: J Materials Design and Applications 2019; 233: 714–720.

16.

Zhao

Zhang

Gao

, et al. Research on calibration method of walking axle torque test of Combine Harvester. Agric Equip Veh Eng 2019; 57: 63–65.

Dynamic load distribution property of transmission system during rice harvesting process

Abstract

Keywords

Introduction

Drivetrain analysis

Transmission structure and parameters

Transmission system structure

Parameters of the main components of the transmission system

Belt strength calculation

Transmission chain strength calculation

Drive shaft strength calculation

Load testing system construction

Test rice material

Dynamic load test method

Calibration of wireless load test components

Field load test on the combine harvester

Test and analysis of load influencing factors

Walking speed test

Effect test of stubble height

Machine load impact test

Road condition impact test

Tests under different walking speed

Ultimate load test of each component

Harvest ultimate load test

Chassis ultimate load test

Transmission system calibration of combine harvester

Strength evaluation of the transmission chain of combine harvester

Transmission belt strength evaluation of combine harvester

Optimization of the transmission system of combine harvester

Improvement and optimization of the drive belt of combine harvester

Improvement and optimization of the drive chain of combine harvester

Improvement and optimization of the drive shaft of combine harvester

Conclusions

Footnotes

Declaration of conflicting interests

Funding

ORCID iD

Data availability

References